Cast Copper Alloy Sleeve Bearings

CAST COPPER ALLOY SLEEVE BEARINGS

Non-Ferrous Founders' Society CAST COPPER ALLOY SLEEVE BEARINGS

INTRODUCTION TO LUBRICATED BRONZE BEARING DESIGN

HISTORICAL...... 7 WHAT MUST BEARINGS DO? ...... 8 DIFFERENT BEARING CLASSES...... 9 Differences Between Sliding and Rolling Bearings ...... 9 Types of Sliding Bearings ...... 10 Dry Bearings and Boundary Lubricated Bearings...... 10 Mixed-Film Bearings...... 10 Full-Film Bearings...... 10 PRINCIPAL BEARING GEOMETRIES...... 10 BEARING OPERATING CHARACTERISTICS...... 11 Bearing Friction, Wear and Temperature ...... 12 BEARING MATERIAL SELECTION CRITERIA...... 14 CAST BRONZE BEARING MATERIALS ...... 14 OTHER BENEFITS OF BEARING BRONZES ...... 15

SELECTING CAST BEARING BRONZES

BRONZES AND COPPER ALLOYS ...... 17 ECONOMICS AND AVAILABILITY ...... 17 DETAILED DISCUSSION OF SPECIFIC ALLOYS...... 19 Tin Bronzes...... 19 Leaded Tin Bronzes ...... 19 High-Leaded Tin Bronzes...... 19 High-Strength Brasses ...... 20 Aluminum Bronzes...... 20 Silicon Brasses...... 21 Copper Beryllium Alloys...... 21 Lead-Free Bearing Bronzes...... 21 Effect of Other Alloying Elements ...... 21 SUMMARY OF CAST BRONZE CHARACTERISTICS AND USAGE ...... 21

TABLE OF CONTENTS\continued

SELECTION PROCEDURE FOR BEARING MATERIALS ...... 21 Initial Selection of Candidate Materials Based on Past Usage ...... 23 Additional Materials Selection Considerations...... 23 Physical Properties of Bearing Materials...... 25 Compatibility with Working Fluids...... 25

DESIGN OF BOUNDARY LUBRICATED BRONZE BEARINGS

CONDITIONS LEADING TO BOUNDARY LUBRICATION ...... 29 BOUNDARY LUBRICATED WEAR DATA FOR CAST BRONZE BEARINGS ...... 29 LUBRICATED WEAR LIMITS FOR BEARING BRONZES ...... 30 LUBRICATED WEAR RATES FOR BEARING BRONZES ...... 31 BEARING TEMPERATURES...... 31 Bearing PV factors...... 32 DESIGN CONSIDERATIONS ...... 32 Maximum Allowable Bearing Temperatures...... 32 Maximum Allowable Bearing Stress ...... 32 Clearance ...... 33 Shafting...... 33 Surface Finish...... 33 Wall Thickness and Flange Dimensions ...... 33 Methods of Retaining Bearings...... 34 Press or Shrink Fit ...... 34 Keying Methods...... 35 Grease Grooving Patterns...... 35 Straight Axial Supply Groove...... 36 Cicumferential Grooves ...... 36 Recirculating Configurations...... 36 Groove Geometry ...... 36 Grease Consumption...... 37 Regreasing Intervals ...... 38

HYDRODYNAMIC JOURNAL BEARING DESIGN

LOAD CAPACITY OF HYDRODYNAMIC JOURNAL BEARINGS...... 41 THERMAL EFFECTS ...... 43 TRANSITION TO HYDRODYNAMIC LUBRICATION ...... 43 DESIGN CRITERIA FOR JOURNAL BEARINGS...... 44 Bearing Size ...... 44 Minimum Recommended Film Thickness...... 44 Design Clearances ...... 45 Maximum Temperatures and Temperature Rises...... 46 Bearing Stresses and Materials...... 46 Safety Factors...... 47 Lubricants and Filtration ...... 47 Geometry Considerations...... 47 Lubricant Grooving...... 47 Machining Considerations ...... 47

OTHER JOURNAL BEARING DETAILS...... 47 Bearing and Load Rotation...... 47 Geometries to Mitigate Edge Loading due to Shaft Bending...... 48 MATING SURFACES ...... 48 Shafting Materials...... 48 Surface Preparation ...... 49 Surface Roughness Levels...... 49 Hardness of Shafts...... 49 OTHER ISSUES...... 50

COMPUTER PROGRAMS

BACKGROUND TO THE PROGRAMS...... 63 USING THE HYDRODYNAMIC PROGRAM ...... 64 Entering Bearing Data ...... 55 Lubricant Selection ...... 65 Entering Data in Alternate Unit Systems...... 66 CALCULATING BEARING PERFORMANCE...... 66 Clearance...... 68 Air-Cooled Design ...... 58 APPLYING THE HYDRODYNAMIC PROGRAM RESULTS...... 59 USING THE BOUNDARY LUBRICATION PROGRAM ...... 60 Entering Bearing Data...... 60 Friction...... 60 Units...... 60 Calculating Bearing Performance ...... 61 REFERENCES AND APPENDICES REFERENCES ...... 65 APPENDIX A. Quantifying Lubricated Wear ...... 67 Archard-Holm Equation ...... 67 Other Forms of the Archard Equation...... 68 Lubricated Wear under Variable Loading Conditions...... 68 APPENDIX B. Thermal Modeling of the Bearing...... 69 General Heat Transfer Model...... 69 Control of Heat Transfer Mechanisms ...... 70 APPENDIX C. Fluid Viscosity Database...... 71 Adding Additional Fluids to the Database ...... 71 Viscosity Conversions...... 71 ACKNOWLEDGMENTS We wish to thank Joseph L. Tevaarwerk for editorial production of this publication. We wish to thank the following for providing photography used in this publication. Bunting Bearings Corporation Magnolia Metal Corporation

This Handbook has been prepared for the use of engineers, designers and purchasing managers involved in the selection, design, manufacture or use of copper alloy bearings. It has been compiled from information supplied by testing, research, manufacturing, standards, and consulting organizations that Copper Development Association Inc. believes to be competent sources for such data. However, CDA assumes no responsibility or liability of any kind in connection with the Handbook or its use by any person or organization and makes no representations or warranties of any kind thereby.

Published 1997 by Copper Development Association Inc., 260 Madison Avenue, New York, NY 10016

A1063-00/97

LIST OF SYMBOLS AND UNITS USED Symbol Explanation Units A apparent area of surface contact m2 2 AB outer surface area of bearing housing m 2 Aj inside bearing area (πBD) m 2 Ar real area of surface contact m B bearing width m

C radial clearance (Db-Dj)/2 m

CD diametral clearance m

Ceff effective radial clearance at θeff m d diametral wear m d* critical wear depth m

D, Db bearing diameter m

Dj journal diameter m E modulus of elasticity N/m2 e bearing eccentricity m F bearing radial load N f friction coefficient -

Fn applied radial load N H hardness N/m2 2 hc convection heat transfer coefficient W/m • °K

hmin minimum film thickness m

h2min recommended minimum film thickness m 2 hr radiant heat transfer coefficient W/m • °K 4 K wear coefficient -

LIST OF SYMBOLS AND UNITS USED\continued Symbol Explanation Units k conduction heat transfer coefficient W/m • °K

ka heat transfer coefficient to ambient W/m • °K L total sliding distance m N journal speed rps

Ne equivalent journal rotation rps

Nj original journal rotation rps

Nb bearing rotation rps

NF load vector rotation. rps

NT transition journal speed rps 2 P nominal bearing stress Fn/B • D N/m

Pta heat flow to ambient W R journal radius, D/2 m Ra arithmetic average of surface roughness µm S Sommerfeld number - T total operational time s t sliding time s

te equivalent sliding time t

ts total sliding time s

Tregrease regreasing interval h U sliding velocity m/s V wear volume m3 3 Vf bearing grease consumption rate cm /h 3 VL total grease volume in bearing m w power loss W α thermal expansion coefficient m/m • °K ε eccentricity ratio -

εT transition eccentricity ratio - η dynamic fluid viscosity Pa • s

ηeff effective viscosity at θeff Pa • s Λ bearing loading factor - ν kinematic fluid viscosity m2/s

θamb ambient temperature °C

θB bearing temperature °C

θeff effective bearing temperature °C

θmax maximum lubricant temperature °C ρ material density kg/m3

σy material yield strength Pa 5

Introduction to Lubricated Bronze Bearing Design I. INTRODUCTION

When two surfaces are in important among these are the design, including the many alloys relative sliding motion, a normal geometry and motion of the bear- now being successfully applied, and load may be transmitted between ing surfaces, the nature and consis- to introduce Hydro and Bound, them. Such a system is referred to tency of the fluid employed, and CDA's bearing design software. as a sliding bearing. Under the continual supply of this fluid. Together, this text and the software nominally dry conditions, this The benefit of having the load can assist designers in choosing the transmitted load is supported by supported by fluid pressure is very optimum material and dimensions small irregularities on the bearing great indeed. It reduces wear and for anticipated service conditions. surfaces, called asperities. These friction by several orders of magni- asperities are in direct contact. tude. Unfortunately we cannot Historical Under limited conditions of load ensure that the correct conditions The problem of transporting and speed, such a bearing will work exist at all times. Contact between heavy loads was significantly quite well and experience moderate two bearing surfaces will invari- reduced by the invention of the wear. Materials such as high-leaded ably occur at some point during wheel. The invention of the wheel tin bronzes have self-lubricating operation. When contact occurs, was significant, but the bearing sur- properties that permit fairly high behavior of the bearing material is face that supported the wheel must loads to be supported this way. very important in the reliable oper- be regarded with equal importance. If a lubricant in the form of a ation of the bearing. Hence, besides How and when the bearing and fluid or grease is present between the bearing geometry, bearing kine- wheel were invented is not known, the sliding surfaces, a pressure may matics and lubricant, the bearing but it may be assumed that once develop such that a significant por- materials are of paramount impor- invented, their use spread fairly tion of the load is transmitted by tance. It is thus not difficult to see rapidly. It is also fairly certain that fluid pressure rather than metal to that the design of lubricated sliding soon it was realized that different metal contact. When all the load is bearings can be quite complex. types of wood gave different useful supported by fluid pressure, we Complexity should not, however, lives and that some materials speak of a full-film or hydro- be allowed to act as a deterrent to reduced the amount of work dynamic bearing. In a truly hydro- tackling a bearing design because required for hauling loads. It was dynamic bearing, the two surfaces the rewards for a good design are probably also soon realized that are thus not in contact at all. Since very rich indeed. Well designed smaller shaft bores gave lower fluids are easily sheared without bearings often have very long life friction but also had reduced life. damage, and generally have a expectancies and are very low cost Today we would quickly under- lower shear resistance than most from an operational and mainte- stand these phenomena through the solids, relative motion between the nance point of view. One of the application of our vast knowledge two bearing surfaces takes place in first tilting-pad hydrodynamic base. In antiquity these bearing the fluid film. This results in a bearings ever designed and put in principles would have been deter- lowering of friction and a reduction service in this country is still oper- mined through experimentation, or elimination of wear. To achieve ating today, after 80 years of almost just like the reduction in effort hydrodynamic or full-film lubrica- continuous service. See Snow1. would have been noticed when wet tion in bearings requires special The intent of this publication is conditions prevailed, or when attention to several key variables to describe the principles of animal greases were applied to that influence film formation. Most lubricated bronze journal bearing

7 the bearings. It was probably a light, fostering an understanding equation in 1883 for estimating the combination of increased life and that still serves us quite well today. power loss in a journal bearing, reduced effort that made the lubri- The real breakthrough in knowing the viscosity, diameter, cation of wheel axles a regular experimental work, leading directly speed and load. He assumed con- practice. to the development of the hydrody- stant viscosity and a shaft centered Gradual improvements namic theory, should be credited to in the bearing. Reynolds, in undoubtedly were made all the time Beauchamp Tower. See Cameron2. England, formulated the now to these bearing systems. Materials Discussing his railroad journal famous Reynolds equation in 1886, such as bronze and later, iron, bearing experiments, Tower indicated: and actually proceeded to solve it, would have increased life, but also "...interesting discovery when oil using the results obtained by Tower increased the attention to be paid to bath experiments were nearing in later experiments as an experi- greasing these bearings because of completion. The bearing seized mental verification of his methods. the relative preciousness of these under increased load (625 psi), and Around the beginning of the 20th materials. In addition to animal and the brass bearing was removed. century, Stribeck and Sommerfeld vegetable fats, the use of pitch-tar, While the bearing was removed, a in Germany advanced the hydrody- tree resins, and beeswax for bearing 1/2-inch hole was drilled for a namic lubrication theory toward lubricants would have developed lubricator fitting through the cast engineering application. Stribeck over time. iron bearing cap and the brass bear- established the minimum friction How and why bearings worked, ing. The bearing was reassembled point and experimentally devel- and why the introduction of lubri- and put in an oil bath lubrication. oped the "Stribeck Curve." cants in them lowered the friction Oil was observed running out of Sommerfeld produced an analytical remained much of a mystery until the hole when the bearing was solution to the Reynolds equation the late 18th and into the 19th cen- operated, making a mess. Attempts for journal bearings, making possible tury. Most of the developments in were made to plug the hole first the development of a Stribeck bearings during this time were dri- with a cork and then with a wooden curve from theory. ven by the wheel bearing problems plug. Each one was forced out, that occurred in the railways. The indicating high pressure. A 200-psi What Must Bearings Do? railways of the time were searching pressure gage was screwed into the Bearings are multifunctional for ways to increase the load and hole and the gage read the devices. In order to operate effi- speed capacity of cars and locomo- maximum gage capacity of 200 psi ciently and provide long service tives. The limit was always set by when the bearing was operated. The life, bearings often have to satisfy the problems encountered by estimated bearing pressure was 100 several requirements simultaneously. bearing failures. psi." These include: Matters were not helped along This indicated to Tower that ■ Position and support a shaft by the almost simultaneous intro- the shaft was floating on a film of or journal and permit motion duction of plentiful, but somewhat oil whose pressure was more than with minimum energy deficient, mineral oils. Oiliness, the double the mean bearing pressure. consumption; ability to prevent wear under metal- Beauchamp Tower's publications ■ Support a fixed load and be to-metal contact conditions, is on his landmark bearing experiments able to withstand occasional almost absent in mineral oils, com- stimulated Osborne Reynolds to deve- shock loads; pared to the vegetable and animal lop the "physical wedge concept" for ■ Run quietly and suppress oils and fats used up to that time. It hydrodynamic lubrication in 1888. externally generated vibrations; was the development of full-film Once the law of viscous flow ■ Act as a guide to support reci lubrication as we call it today, that was well known, and it was recog- procating or oscillating motion; gave the breakthrough into bear- nized that bearings developed their ■ Withstand temperature ings capable of high loads and high own pressurized lubricant films, excursions; speeds. Advances in both experi- the possibility of mathematical ■ Accommodate some degree of mental techniques and mathematical prediction of bearing behavior was shaft misalignment; analysis brought a lot of detail to realized. Petroff developed an

■ Accommodate dirt particles trapped in the lubricant; ■ Resist corrosion — under normal service conditions as well as during storage or extended down-time; and ■ Provide easy maintenance. As with all engineering endeav- ors, selection of the type of bearing and the bearing material usually involves some compromise among Figure 1.1. Load Carried by Rolling (I) and Sliding (r) Motion. the often competing design requirements. Bronze bearing alloys cannot satisfy all needs at all times, but they do offer the broadest range of properties among today's sleeve bearing materials and can perform well under a very wide range of operating conditions.

Different Bearing Classes When we speak of bearings, we mean a device that can transmit a load from one surface to another, when these surfaces experience a relative motion. This may be achieved by either some form of rolling motion or by some form of sliding action, see Figure 1.1. Based on this fundamental difference we divide bearings into two broad classes as either: Rolling-Element Bearings (REB) or Sliding-Surface Bearings (SSB). These two classes are then further subdivided according to different forms of surface geometry and different types of surface interactions, see Figure 1.2. Figure 1.2. Bearing Classes and Different Types within Each Class. Differences Between Sliding and Table 1 Rolling Bearings As may be observed in on because they can be custom-made page 10, there are so many inherent economically for an almost limit- There are important differences advantages in sliding bearings that less variety of applications— between the two classes of bearings one should make every available including and exceeding all those that are worth noting. Many of effort to use them. Sliding bearings to which rolling-element bearings these differences will favor the use are used because they are simple, can lay claim. of one type of bearing over the other. reliable, compact, durable, and

9 cant can be maintained and provided Sliding Bearings Rolling Element Bearings in sufficient quantity for successful Need to be purposefully designed Off-the-shelf finished bearings operation. Capillary action within and manufactured. available. the network of voids helps supply Bearing, per se, usually less costly. Lower overall cost for larger bearings. the surface with lubricant. Addition- Overall cost is less for small bearing Small bearings may have higher al lubricant may be available to the systems. relative cost. bearing in the form of storage in Service life not predictable, but can be Statistically predictable finite life. felt reservoirs. capable of providing unlimited life. Offer more design flexibility and versatility Limited by available standard sizes, and broader spectrum of materials. configurations and materials. Full-Film Bearings In these types of bearings the More tolerant of abuse. May require Very sensitive to abuse. Minimal periodic maintenance. maintenance, short of replacement. lubricant is applied to the working surface from an external source. Can provide high levels of beneficial Minimal damping characteristics. Lubricants typically used are oils damping. or greases. They may also be other Sometimes higher friction. Wear, under Finite, generally low friction fluids such as water, process fluids and wear. favorable conditions, can approach zero. or even gases. Full-film bearings Little outside help available. Lots of outside help available fall into one of four categories: (suppliers). ■ Hydrodynamic, where the shaft or bearing rotation gener- Table 1. Differences between Sliding and Rolling Bearings. ates a thin load carrying lubricant film, ■ Squeeze-film bearings, where The use of rolling element disulfide (MoS2) and PTFE. A very bearings in every conceivable successful line of dry bearings uses the normal motion of the application is based more on the a sintered bronze material on a surfaces generates a thin lack of suitable design skills in steel backing. The voids in the lubricant film ( connecting rods sliding bearing design, rather than sintered material are then filled in internal combustion engines), an inherent superiority of rolling with a mix of lead and PTFE. ■ Hydrostatic, where lubricant is element bearings per se. Boundary lubrication exits in supplied under very high liquid (or grease) lubricated bear- pressure sufficient to separate the metal surfaces, and Types of Sliding Bearings ings when there is no load carrying film and the lubricant serves mainly ■ Hybrid bearings, where both Within the sliding bearing to keep friction fairly low. Cast hydrostatic and hydrodynamic category there are various types of bronze bearings are used extensively features are used. bearings broadly categorized as for boundary lubricated applica- follows: tions. We will return to this topic Principal Bearing Geometries ■ Dry bearings and boundary later. lubricated bearings, Sliding bearings may be further ■ Mixed-film bearings, and categorized by the direction of load Mixed-Film Bearings support that they provide. Broadly ■ Full-film bearings. Oil-impregnated porous metal speaking this would be the journal bearings are the most commonly bearing and the thrust bearing, see Dry Bearings and Boundary used lubricated bearings where Figure 1.3. Lubricated Bearings insufficient film exists to carry all The journal bearing, also In the dry bearings, the lubri- the load. Only under very favorable known as a sleeve bearing, a bush- cant is provided by solid particles conditions will the complete load ing or a radial bearing, can support contained within the bulk material. be supported by the fluid film a radial load on a shaft. This radial Besides lead, typical solid lubri- developed. These bearings contain load may act in any orientation, and cants include graphite, molybdenum voids within which a liquid lubri- may vary with time. Each of the

10 different lubrication modes with a thrust bearing requires of the effort needed to move mentioned above is possible with surface features that need to be the surfaces relative to a journal bearing. In its simplest machined into the thrust washer each other, form, a journal bearing consists of surface. They may consist of sim- ■ Wear coefficient, a measure a shaft that fits through a clearance ple grooves or dimples, or very of the amount of material hole in a plate. This simple geometry sophisticated shapes such as loss during the sliding and a is such that a full-film lubrication wedges and spiral groove patterns. determining factor in bearing of the shaft and bearing may be Because of the additional complexity durability, and possible under certain circumstances. in designing hydrodynamic thrust ■ Local bearing temperature, The clearance space between the bearings, we will limit this discus- a measure of the likelihood shaft and bearing form a natural sion to journal bearing designs. of sudden failure that may wedge, and no complex machining be experienced in the system. techniques are needed to make Bearing Operating Variations of these design a simple hydrodynamic journal parameters are shown as a function bearing. Characteristics of a so-called hydrodynamic vari- Sliding bearings can operate in A thrust bearing is intended to able that contains velocity, viscosity three fundamental modes, and it is support an axial thrust load on a and load as important variables. In important to understand that, for a shaft. It may consist of a simple essence, the hydrodynamic variable given set of system requirements, flat washer of suitable material that determines how a load is supported in a sliding bearing. Hence, the load transmitted between two bearing surfaces may be carried entirely by a continuous hydrodynamic lubricant film, entirely by the surface roughness peaks on the two opposing surfaces, or by a combination of the two. Thus, depending on the method by which the load is carried between two surfaces, we may define three lubrication and corresponding wear regimes as: ■ Boundary lubrication. Large- scale asperity contact occurs. Load is carried by asperities Figure 1.3. Machine Member with Journal Bearings and a Thrust Bearing. and wear is moderate to heavy, depending on the nature of the surfaces and the materials their operating modes can be con- rides against a thrust collar on the involved. trolled by design. A good basis shaft. This simple form of thrust ■ Mixed-film lubrication. A bearing will only be able to operate for the understanding of the fundamental phenomena in sliding lubricant film is present, but under boundary lubricating condi- intermittent contact occurs. tions because no natural wedge bearings is the modified Stribeck diagram shown in Figure 1.4. This Load is carried partially by action takes place. Hence, similar fluid pressure and partially by to journal bearings, thrust bearings diagram presents three lubricated bearing performance parameters asperity contact. Wear is benefit greatly from the boundary moderate to light, depending lubrication characteristics of that are critical in slider bearing on the chemical activity of cast bronzes. design. They are: ■ Friction coefficient, a measure the lubricant and bearing To obtain full-film lubrication

11 from low to high hydrodynamics, and then experiences a rather sudden drop in the "mixed-film lubrication" region. The lowest friction occurs at the "transition point," and after this the friction rises slowly. The explanation for this friction behavior lies in the fact that friction at low hydrodynamics is almost completely governed by surface interaction. At high hydrodynamics, it is almost exclusively fluid friction (dashed line). Asperity friction may be due to adhesive or abrasive contact. The friction level depends on the bearing material and its chemical interaction with the lubricant. Friction level is nearly velocity independent. To achieve low friction in the boundary lubricated region is thus a materials selection issue. In the fluid friction region the nature of the bearing surfaces is not relevant because there is no asperity contact. The direct interac- Figure 1.4. Modified Stribeck Curve for Lubricated Sliding Surfaces. tion properties of the bearing and shaft materials are not important materials and on the contact- straints may preclude operation in here; however, the material must ing surfaces. this domain and the bearing may have sufficient plastic flow ■ Load carried by Hydrodynamic have to operate for considerable strength (indicated by hardness) action only. When conditions periods of time in the boundary and fatigue resistance to avoid of full hydrodynamics exist, lubricated regime. gross dimensional changes due to the surfaces are completely high fluid pressures. separated by a lubricant film, and the load is carried by fluid Bearing Wear (upper half of pressure only. Wear in this Bearing Friction, Wear and Figure 1.4.) decreases drastically as regime is very low to non- Temperature we move from low to high hydro- existent. Examination of the behavior dynamics, showing a rather sudden reduction in the "mixed film lubri- Transition from boundary to of bearing friction, wear and tem- cation" region. This sudden drop hydrodynamic lubrication occurs perature as a function of the bearing occurs for the same reason that the when the mean film thickness sep- operating regimes gives us some friction drops, namely the reduc- arating the surfaces is greater than important insight into bearing tion in asperity interaction due to the local system roughness. Low materials selection and overall load support by the lubricant film. roughness tends to promote early bearing design. transition to hydrodynamics. It is Bearing wear can occur due to Bearing Friction (solid line in several mechanisms. In sliding always the desired goal in bearing the lower half of Figure 1.4.) remains design to work to the right of this bearings, we typically encounter almost constant as we start to move wear due to adhesion, two-body transition. However several con-

12 abrasion, or three-body abrasion. combated. If the source of particles the propensity for asperities to Adhesive wear occurs when inter- is external, we may employ lubri- weld together increases rapidly in acting materials form solid bonds cant filters and seals to reduce the the boundary lubricated region. at the points of contact. concentration in the bearing. Third- Depending on the materials, a Sliding of the surfaces relative to body abrasion due to self-generat- sufficiently high temperature may each other causes the breakage of ed wear particles can be reduced be reached to cause scuffing dam- these bonds and results in removal of by increasing the lubricant flushing age. Scuffing damage may be cata- surface material. Two-body abrasion action. We could also use harder, strophic to a bearing operation in occurs when metal surfaces are in more wear resistant bearing materials. the form of galling and seizure. direct contact and where asperities To avoid extensive bearing damage Key factors that tend to keep from a hard surface penetrate into the due to very large hard abrasive par- the peak bearing temperatures low other softer surface. Relative motion ticles, an accommodation mecha- are low friction (low adhesion), between the two now causes material nism is incorporated by providing a early transition, good heat conduc- removal of the softer surface. Both fairly thick soft layer. Large, hard tion and low, local normal stress. adhesive wear and two-body abra- particles embed themselves into As far as material selection is con- sive wear occur in sliding bearings this layer. This however can lead cerned, we should select soft mate- when the surface separation is less to the two-body abrasion of the rials with low friction and high than the mean operating roughness. harder (normally shaft) material. conductivity. Copper based bearing Adhesive bearing wear can be materials possess several of these Bearing Temperature (dotted line combated by selecting bearing mate- key requirements. in Figure 1.4.), which results from rials with low solubility into each the product of friction and sliding From the foregoing discussion other. Materials such as indium, velocity, is low at low sliding of lubrication regimes we should cadmium, lead, tin, silver and bis- speed, reaches a peak, and then note that a very important consid- muth have low solubility with steel reduces due to a drop in friction. eration of the bearing design is the and are thus excellent choices as Further increases in speed pro- bearing material itself. This is bearing materials. Note all the mate- duces a minimum temperature, and where the various bronze materials rials mentioned are very soft. It is then it starts to rise again. The ini- are providing us with an excellent also possible to add substances to a tial rise in temperature is almost choice. lubricant to contaminate the likely exclusively the result of solid sur- bonding sites. Readers interested in a more in- face friction, while at high hydro- depth explanation of asperity inter- Two-body abrasion can be min- dynamics it is completely due to action and the impact of materials imized by selecting harder bearing 3 fluid friction. may wish to consult Rabinowicz . materials or by reducing the surface There are several important An excellent treatise on the differ- roughness on the harder of the two observations that can be made ent wear modes in bearings is surfaces. Increasing hardness of the 4 about the bearing temperature. given by Lansdown . bearing material may be detrimen- The initial rapid rise is the result of tal to some other important aspect high friction coupled with the of the bearing. velocity difference between the Three-body abrasion is caused two surfaces. Remember, this is the by hard particles in the lubricant. boundary lubrication regime, and They may originate from sources almost all the bearing load is car- outside the bearing such as dust ried by direct asperity contact. The and sand, or they may be generated friction in this regime is due to within the bearing system itself due adhesive and abrasive asperity con- to adhesive or two-body abrasion. tact. It is well known that increas- There are several ways in ing temperature tends to promote which three-body abrasion can be adhesion between materials. Hence

13 Bearing Material bearings are run against hard- ■ High creep strength. Some Selection Criteria ened steel shafts. bearings must withstand ■ The ability to conform to elevated temperatures. In such Alloy selection is obviously non-uniformities in the shaft cases, creep and stress-rupture very important in the design of surface and to shaft properties have to be taken bearings and especially when oper- misalignment. Low strength into account. ating in the boundary lubricated or relative softness are some- ■ High fatigue strength. Some regime. While the choice of alloy times advantageous in a bearings are subject to is less critical for full-film bear- bearing alloy. Not all shafts repeated pounding, others to ings, it does influence performance are turned to precision toler- pulsating or vibratory stresses. during the non-hydrodynamic ances or finishes, and service High fatigue strength generally conditions at start-up or upset oper- conditions can provoke mis- accompanies high tensile ation brought about by shock alignment; a softer alloy can strength, but there are loads, dirt intrusion, temperature tolerate such defects. exceptions to this rule. surges and intermittent lubricant ■ Good corrosion resistance. ■ High thermal conductivity. flow. Certainly, the more precisely Pump bearings are routinely Heat is damaging to most a bearing's operating conditions exposed to sea water, sour bearings and lubricants, there- can be defined, the simpler and mine waters, acids and any fore the ability to conduct more obvious will be the choice number of corrosive industrial frictional heat away from the of bearing materials. process streams. All bearings bearing surface results in Selection of the specific bear- occasionally get wet, and even lower operating temperatures. ing alloy requires a very careful some constituents of lubricants ■ A uniform metallurgical consideration of the application are or can become aggressive. structure. Sound, non-porous and the characteristics of the mate- It is therefore important that surfaces are needed to promote rial itself. There is no single best bearing materials exhibit and maintain a stable lubricant bearing material. The following are sufficient corrosion resistance, film. considered to be the attributes of a both in service and during ■ Compatible thermal "good" bearing material: storage. expansion. The thermal ■ A low coefficient of friction ■ High compressive strength expansion of the bearing must against steel. The need for this combined, where needed, with be known in order to establish property is self-evident, good impact strength. proper clearances at the bear- however good inherent Strength obviously determines ing's operating temperature. frictional properties are partic- the bearing's load-carrying Some of these attributes are ularly important in cases of capacity, while impact unreliable lubrication, i.e., obviously more important than behavior frequently dictates others, and only a few may be boundary- or mixed-film the permissible types of conditions. In such cases, the critical in particular installations. service. Alloys with low All, however, influence bearing metal itself must be relied impact strength must be upon to provide adequate anti- performance to some degree, and adequately supported in none should be ignored. friction behavior until lubrica- suitable housings. tion is restored. ■ Low shear strength. This Cast Bronze ■ High wear resistance. This property is characteristic of property is different from, but Bearing Materials certain leaded alloys, whose related to strength, hardness, Literally hundreds of different ability to smear across the friction and a number of other bearing alloys are available for use shaft surface provides a form properties. As with other prop- with steel mating surfaces. Along of "internal lubrication." This erties, wear resistance is most with sintered and wrought bronzes, is highly beneficial when important when lubrication aluminum and zinc alloys and a normal lubrication is intermit- momentarily fails. Wear resis- variety of polymers and composites, tent or unreliable. tance is also essential when these materials make up many of

14 the sliding bearings in Load Capacity Maximum Conformability Hardness use today. For additional and Fatigue Operating and Resistance to and Wear information, the reader Bearing Alloy Families Resistance Temperature Embeddability Seizure Resistance is referred to an excel- Leaded Red Brasses Moderate Moderate Low Low Good lent detailed review Tin Bronzes High High Moderate Moderate High by Pratt5 and a more general description High-Leaded Moderate / High Good Good High Tin Bronzes High by Glaeser6. Many bearing Aluminum Bronzes Very High Very High Poor Moderate Very High materials offer promis- High-Strength Brasses Moderate Moderate Poor Moderate High ing individual proper- High Poor Poor ties (light weight, low Copper Beryllium Alloys High Very High friction) to meet partic- Leaded Coppers Moderate High Very Good Very Good Moderate ular service require- Tin-Based Babbitts Moderate Moderate Excellent Excellent Low ments, but the cast bearing bronzes offer Lead-Based Babbitts Moderate Moderate Excellent Excellent Low the broadest range of Aluminum Low-Tin High High Good Low Moderate applicability. Their or Lead Alloys favorable combinations Aluminum High-Tin Moderate High Good Moderate Moderate of mechanical and or Lead Alloys physical properties allow the designer to Table 1.2. Comparison of Important Bearing Selection Criteria for Commonly Used Materials. optimize a bearing without backing, design without compromising Other Benefits of ■ no need to rely upon specialist needed characteristics unnecessarily. Bearing Bronzes manufacturers and suppliers, Table 1.2. from CDA7 compares Besides the superior strength and manufacture can usually many of the tribologically impor- and wear properties of bearing be in-house, tant criteria for a range of common- bronzes, there are additional bene- ■ easy to fine tune bearings. If a ly used materials. fits when using these materials. characteristic of a specific Note the broad range of bear- Some of these are: bearing is found deficient, a ing bronzes to choose from, and ■ bearing bronze materials are review can be made of the the superior load and wear perfor- readily available, important material properties mance of the copper based materials. ■ all grades can be produced in needed and a selection made These latter two properties become almost any size in large or from available materials to extremely important when less small quantities, provide improved properties. than ideal lubrication conditions ■ normal machine shop facilities The large selection of cast exist. Chapter 2 contains descrip- are adequate for preparation, copper-based bearing alloys tions of the various bearing bronzes ■ substantial structural strength provides a wide choice of in greater detail. permits the use of bearing engineering properties, so that bronzes as an integral part of incremental adjustments can the structure. This inherent be made. strength also permits their use

15 Properties and Selection of Cast Bearing Bronzes SELECTING CAST BEARING BRONZES

A very important aspect of selection of materials. Therefore, improved mechanical properties. sliding bearing design is the selec- the selection of a material also A further variation is centrifu- tion of the actual bearing materials. requires a review of the economic gal casting in which the mold is Literally hundreds of different factors. Overall costs are frequently rotated during casting. Any impuri- bearing alloys are available for use the deciding factor in selection of a ties which are present are usually with steel mating surfaces. Selection bearing material. Some materials less dense and are therefore sepa- of a specific bearing alloy requires cost more than others. Bearing rated towards the center of the very careful consideration of the costs are, to a certain extent, sub- mold where they can be removed, application and the material char- jective. The economic assessment and as a consequence a cleaner acteristics themselves. There is no should also include factors such as alloy is produced. With centrifugal single best bearing material. availability, ease of manufacture, casting methods, however, segre- Technically, an assessment must be replacement cost, the degree of gation of high lead content alloys made of component and system precision required in its manufac- can occur. Continuously cast alloys requirements to allow placing a ture and installation, spare parts are also of high quality, with prop- priority on the material properties cost, and ability of making incre- erties comparable to alloys cast by for a given application. mental adjustments to the bearing. centrifugal casting. While bronze bearing alloys The cast bearing bronzes are are certainly not the least expen- available in a great variety of sizes Bronzes and Copper Alloys sive engineering materials, they are and shapes. Particularly suitable Many bearing materials offer easily competitive in terms of for the manufacture of bearings are promising individual properties performance and life-cycle costs. the following: Finally, and not incidentally, (light weight, low friction) to meet Tin and leaded tin bronzes: bronze sleeve bearings offer the particular service requirements, but ■ As finished bearings in sizes distinct advantage that virtually all the cast bearing bronzes offer the ranging from 3 mm to 152 mm grades can be produced in almost 5 broadest range of applicability. ( / in to 6 in ) in diameter, up any size in large or small quantities. 16 Their favorable combinations of to 230 mm (9 in) long; Cast bronze bearing alloys are mechanical and physical properties ■ As solid bars ranging from available in a large variety of stan- allow the designer to optimize a 10 mm to 260 mm (3/ in to dard and custom shapes and sizes. 8 bearing design without compro- 10 in) in diameter, up to Bearing blanks are cast as cylindri- mising needed characteristics 2700 mm (105 in) long; and, cal shapes using the one of several unnecessarily. The large selection ■ As cored bars 25 mm to of cast copper-based bearing alloys techniques available. They may be 300 mm (1 in to 12 in) in diam- provides a wide choice of engi- sand cast, chill cast, continuous eter, in various bore sizes, also neering properties. Table 2.1, com- cast or centrifugally cast. The rate up to 2700 mm (105 in) long. piled from data presented in CDA8, of cooling of the molten alloy Aluminum bronzes: gives the composition of the most influences the grain size of the ■ As solid bars 13mm to 200 mm important cast bearing bronzes in solidified material, and slow cool- (0.5 in to 8.0 in) in diameter, use today. ing usually gives a coarser structure with reduced mechanical from 3,000 mm to 3,700 mm

Economics and Availability properties. Chill casting can be (120 in to 144 in) long; used instead of sand casting to get ■ As cored bars 30 mm to 260 mm As with all engineering, eco- (1.25 in to 10 in) in diameter, in nomic factors play heavily in the

17 Designation Nominal Composition, Wt% UNS SAE Cu Sn Pb Zn Fe Ni P Al Other Leaded Red Brass C83600 40 85 5 5 5 _ Tin Bronzes C90300 620 88 8 4 _ C90500 62 88 10 - 2 - - - - C90700 65 88 11 - - - - - Leaded Tin Bronzes C92200 622 88 6 1.5 4.5 _ _ _ _ C92300 - 87.3 8 0.7 4 - - - - C92700 63 88 10 2 - - - - - High-Leaded Tin Bronzes C93200 660 83 7 7 3 _ C93400 - 84 8 8 - - - - - C93500 66 85 5 9 1 - - - - C93600 M-64 80 7 12 1 - - - C93700 64 80 10 10 - - - - - C93800 67 78 7 15 - - - - - C94100 - 73 5.5 20 1 - - - - C94300 (AMS-4840) 68 5 25 .8 - 1 - - C94500 - 72 7 19 1.2 - 1 - - High-Strength Brasses C86300 430B 63 _ _ 25 3 6 3 Mn C86400 59 - 1 40 - - - - Aluminum Bronzes C95300 68b 89 1 10 C95300HT - 89 - - - 1 - 10 - C95400 - 85 - - - 4 - 11 - C95400 HT - 85 - - - 4 - 11 - C95500 - 81 - - - 4 4 _ 11 - C95500 HT - 81 - - 4 4 11 - C95520 - 75 - - - 5 5 11 3 Mn C95800 79 - - 4 4.5 9 1 Mn Silicon Brasses C87600 _ 90 _ _ 5.5 _ 4.5 Si C87900 - 65 - - 34 - - - 1 Si Copper Beryllium Alloys C82800 96.6 ------2.6Be/ .5Co/ .25Si Leaded Copper C98820 484 Rem _ 42 _ _ _ _ _ Lead-Free Bronze C89320 - 89 6 - - - - - 5 Bi

Table 2.1. Nominal Compositions of the Most Common Cast Bearing Bronzes.

18 various bore sizes, up to greater the proportion of Cu31Sn8 especially advantageous where 3,700 mm (144 in) long; will be present. The hard inter- high-speed volume production and, metallic compound imparts high and/or complex shapes demand ■ As rectangles, 6 mm x 26 mm wear resistance, but it also causes a relatively large amount of (0.25 in x 1.0 in) to 75 mm the alloys to be somewhat abrasive, machining. (3 in) square, up to 3,700 mm and as a result, the tin bronzes (144 in) long. should be used against hardened High-Leaded Tin Bronzes steel shafts (minimum hardness Custom castings are available This class of alloys consists of 300-400 HB). The addition of zinc in most bearing materials. They the most commonly used bearing to tin bronzes displaces some of range in size from miniatures only materials. They rank somewhat the tin and increases the presence 6 mm (0.25 in) in diameter to large lower than the unleaded or leaded of a strong delta phase. Zinc is rings up to 3 m (10 ft) across. copper-tin alloys described above therefore added as a strengthening in load-carrying capacity, however Detailed Discussion of agent, as in alloys C90300 and they adequately satisfy require- C90500. Specific Alloys ments for bearings operating under The tin bronzes are hard and moderate loads and medium-to- The broad classification of cast strong, and have good corrosion high speeds—the bulk of bearing bearing bronzes in Table 2.1 are resistance, especially against sea- applications. along composition lines. Different water. They are wear resistant and Lead has very low solubility in levels of alloying elements impart withstand pounding well. The tin liquid copper and is almost entirely characteristics to the individual bronzes can be classified as being insoluble in solid copper or its alloys. materials that may be used in the only moderately machinable. Further, lead melts at a relatively materials selection. The following They can be turned, bored and low temperature (327 C, 621 F), is a fairly detailed discussion of the reamed, but are difficult to broach. and when bearings are cast, the specific alloys. The alloys work well with lead in the alloy freezes last, form- grease lubricants. They can func- ing irregular globules between Tin Bronzes tion as boundary-type bearings boundaries of the copper-tin alloy. Tin increases the strength of due to their ability to form polar The lead content in these alloys, copper alloys. Unless it is present compounds with small traces of nominally between 7% and 15%, in high concentrations, it has only lubricants, thus stabilizing (holding is sufficiently high to improve their a small effect on thermal conduc- on to) extremely thin organic bearing properties. The islands of tivity, compared with most other layers and thereby reducing the free lead tend to smear over the alloying elements. Copper-tin chance for metal-to-metal contact. bearing and shaft surfaces acting, bearing alloys, traditionally called Lacking lead, however, the tin in effect, like a built-in lubricant; tin bronzes, also exhibit low fric- bronzes require adequate and this can prevent seizing in the tion coefficients against steel. The reliable lubrication. event of an interruption in the combination of these properties lubricant supply. These alloys are causes the surface temperatures of Leaded Tin Bronzes therefore especially recommended the tin bronzes to tend to remain This group of tin bronzes for boundary-lubricated and lower than other bearing alloys. contains between one and two per- mixed-film applications. The Since heat adversely affects all cent lead, chiefly to improve their alloys can be run against unhard- bearings, this is seen as a distinct machinability. Their lead content is ened shafts. Like the lower-leaded advantage. too small to materially affect their bronzes described above, these All of the alloys in this series bearing properties, and they can be alloys are free-cutting. contain more than 5% tin and thought of as free-cutting versions Since lead is not in solid solu- contain the hard intermetallic of the tin bronzes described above. tion, its effect on thermal conduc- compound Cu31Sn8 in their It therefore follows that they are tivity is much lower than its high microstructure. The more tin used in applications similar to proportion of the microstructure (C90300 < C90500 < C90700), the their lead-free counterparts and are would suggest. High-leaded tin

19 bronzes therefore retain much of good pounding resistance. they have little capacity to embed the favorable thermal conductivity Alloy C93700 exhibits good trapped particles and therefore do of their unleaded counterparts. corrosion resistance against sulfuric not tolerate dirty lubricants. Lead does, however, reduce the acid (in limited concentrations), alloys' strength and ductility, there- acid mine waters, mineral waters Aluminum Bronzes fore alloys with very high lead and paper-mill sulfite liquors. It Aluminum bronzes are the contents (C93800 and C94300 at has excellent wear resistance at highest-strength standard copper- nominally 15% and 25% lead, high speed and heavy pressure based bearing alloys. Aluminum respectively) should be used at and, among alloys in this family, is a potent strengthener in these lower operating stresses and in tolerates shock and vibration well. alloys; in general, the higher the situations in which there is little Alloy C93800 can be charac- alloys' aluminum content, the high- chance of impact or shock loading. terized as having fair strength, er their strength. Alloys with alu- good corrosion resistance to sulfu- In a boundary-type bearing, minum contents higher than about ric acid and sour mine waters. It one rule of thumb advises that 9.5% can be heat treated, and this has fair wear resistance but excel- designers choose the softest (low- is noted in the designation "HT" lent antifriction properties. It is est strength) alloy able to support following the alloys' UNS numbers. non-seizing, readily machinable the applied load. The advantage Alloy C95500HT, for example, and can be used where lubrication offered by this method of alloy can attain tensile strengths in is doubtful. selection lies in the fact that soft excess of 800 MPa (115 ksi). alloys such as high-leaded tin Alloy C94100 has moderate One interesting feature of the bronzes are better able to conform load carrying capacity but, like the aluminum bronzes is their high to shaft misalignment or deflection, other members of this family, it elevated temperature strength. The and can accommodate (embed) dirt exhibits excellent antifriction prop- compressive strength of C95400 at particles carried in the lubricant erties. It is especially good for use 260 C (500 F) is the same as that stream, be it oil or grease. Some under boundary and mixed-film of tin bronzes at room temperature. designers follow this rule to the conditions. point of lengthening the bearing to The aluminum bronzes have the highest fatigue strength of all reduce the applied stress, thereby High-Strength Brasses enabling the use of softer materials. bronze bearing materials, and they The high-strength brasses can resist repeated and severe Alloy C93200 (SAE Alloy (sometimes improperly referred to impact loads very well. One 660) is generally considered to be as manganese bronzes) are modifi- important disadvantage is their rel- the workhorse alloy of this series. cations of the familiar 60% Cu- atively poor machinability. They Originally developed as a low-tin 40% Zn yellow brass known as can be turned, bored and reamed, composition during World War II, Muntz Metal. Alloy C86300 con- but like the tin bronzes, they are the alloy remains in wide use tains manganese, aluminum and difficult to broach. today. Major areas of application iron, which raise its tensile strength Like the high tensile brasses, include light to moderate duty to well over 800 MPa (115 ksi). general purpose bearings. most of the aluminum bronzes Alloy C86400 is somewhat lower must be used against steel shafts Alloy C93600 can be thought in strength, but it contains one hardened to greater than 500 HB. of as a modified version of Alloy percent lead to improve its Alloy C95200 is somewhat more C93200; it contains about one- machinability. Despite their high forgiving in this respect. As a third as much zinc and almost tensile strength, the alloys' fatigue class, aluminum bronzes require twice as much lead as the SAE 660 resistance can only be rated as reliable full-film lubrication to alloy. As a result of the alloy's high moderate. These high-tensile brasses prevent metal-to-metal contact and lead content, its machinability and can operate under very high loads possible scoring. Both bearing and anti-seizing properties are consid- and at moderately high speeds; shaft should be machined to surface erably improved. Reduced zinc however, they require hardened, finishes finer than 0.4-0.5 µm (15- content improves corrosion resis- well aligned shafts and reliable 20 µinch) rms. The alloys do not tance somewhat. High hardness lubrication. Being relatively hard, embed dirt well, and generally and strength impart the alloy with

20 require good shaft alignment. and the alloys can be used in less to accommodate special or severe Their thermal conductivity is only demanding applications. Copper service conditions. For example, slightly lower than some of the beryllium alloys are, however, manganese, aluminum and phos- lower strength alloys, and as a quite expensive and their use is phorus can be added in small result, they can be used at moderate therefore usually reserved for amounts to enhance the properties speeds if shaft hardness, shaft applications in which their strength imparted by one or more of the alignment and lubrication are all can be fully exploited. major alloying elements; tin, zinc, well controlled. The bearings require very hard and in some cases, silicon. shafts, machined to close tolerances Manganese increases strength but Silicon Brasses and very precise alignment. They moderately decreases ductility. It require reliable lubrication. Dirt does, however improve an alloy's Silicon imparts good castability embedding properties are poor. high-temperature working proper- to copper-zinc alloys, and also Although very strong, the alloys ties. Aluminum, a potent strength- adds significantly to the alloys' have fair to poor impact resistance, ener in its own right, tends to raise strength. Silicon brasses are not therefore bearings should be the strength of copper-zinc and among the most common bearing supported as required by suitable copper-tin alloys when added in alloys, but they do have good bear- housings. small amounts. Phosphorus is a ing characteristics at moderately deoxidizer and potent strengthener. high speeds. They are quite readily Lead-Free Bearing Bronze It also improves fluidity during machinable, considering they casting. When used to excess, Bismuth may be substituted contain no lead and are higher in phosphorus can give rise to local- for lead in the cast bronze bearing strength than standard tin bearing ized hard spots, a condition which alloys. Silicon brasses require reli- materials. Lead-free bronzes are obviously should be avoided. able, clean lubrication (dirt embed- formulated to mimic the properties ding ability in these relatively hard of the leaded alloys but without the alloys is only rated as fair). The safety concerns of lead. Friction, Summary of Cast Bronze alloys should be used against yield and tensile strength and Characteristics and Usage hardened shafts. corrosion resistance are all compa- A distillation of the major rable to leaded bearing bronzes. bearing characteristics and usage Copper Beryllium Alloys Alloy C89320 is thus very similar of the different cast bronze materi- in performance to C93200. als is shown in Table 2.2. This Copper beryllium alloys can table may be used for an initial be heat treated to attain higher scan for potential materials or strengths than any other copper Effect of Other Alloying Elements when searching for an improved alloy. Heat treatment is fairly The compositions of the bearing material. straightforward: solution annealing alloys described above have been at 780-800 C (1450-1500 F) is standardized and can be ordered Selection Procedure for followed by water quenching and according to their UNS designa- aging at 340 C (650 F). This heat tion or applicable standard specifi- Bearing Materials treatment, in alloy C82800, cations. As in most cast alloys, The selection of a suitable produces tensile strengths at or allowable compositions are given bearing material for a given appli- higher than 1,100 MPa (165 ksi). in ranges to accommodate foundry cation is a rather difficult task Other copper beryllium alloys practice. Except in the case of because, unlike the theoretical have similar or slightly lower copper beryllium alloys and alu- calculations used for film thickness strengths. minum bronzes, minor deviations determinations, a large number of Copper beryllium alloys have from set compositions have only variables are involved. In practice good bearing properties. When minor effects, if any, on mechani- though, we seldom design a com- fully hardened, they withstand cal property. pletely new bearing system. Often extremely high stresses. Heat treat- Modifications to standard there is a similar design, or an ment can be adjusted to produce a compositions are always possible, extension of a given design that we range of mechanical properties, of course. These may be called for can glean some information from for the start of a material selection.

21 Alloy Characteristics and Uses Leaded Red Brass Reasonable strength, excellent thermal conductivity, reasonable corrosion resistance to sea-water and brine, and good machining and casting properties. Lead content ensures pressure tightness. Leaded Red Brass is used as a low-cost bearing material when low loads and low speeds are encountered. Requires good, reliable lubrication and a moderately hard shaft.

Tin Bronzes Hard, strong alloys with good corrosion resistance, especially against seawater. Moderately machinable. As bearing materials, they are wear resistant and resist pounding well. Best for high loads, low speeds. Require good, reliable lubrication and a moderately hard shaft. Higher tin concentrations improve strength, but at the expense of conformability and embeddability.

Leaded Tin Bronzes Lead improves machinability in these tin bronzes but does not materially affect mechanical properties. The alloys are essentially free-cutting versions of the tin bronzes, above, and have similar properties and uses.

High-Leaded Tin Bronzes Most commonly used bearing alloys, found in bearings operating at moderate loads and moderate-to-high speeds. Excellent for boundary-lubricated situations or where lubrication is uncertain. Since lead is insoluble in the solid phases, it is dispersed in the matrix as small isolated globules, acting as small reservoirs of lubricant. Increasing lead concentrations increase conformability and scoring resistance, but at the expense of reduced strength and pounding resistance. Alloy C93200 is considered the workhorse alloy of the series. Alloy C93600 has improved machining and anti-seizing properties. C93800 is noted for its good corrosion resistance against concentrations of sulfuric acid below 78%. Alloy C94100 is especially good under boundary- lubricated conditions.

High-Strength Brasses Alloys with high mechanical strength, good corrosion resistance and favorable castability, and of low relative cost. Can be machined but, with the exception of C86400 and C86700, are less readily machined than leaded compositions. Brasses have typically poor tribological properties, and hence as a bearing material tend to be used in non-critical applications. When used for high-strength bearings, alloys C86300 and C86400 require hardened shafts and reliable lubricant supplies.

Aluminum Bronzes The aluminum bronzes are characterized by high strength and excellent corrosion resistance. Some can be heat treated ("HT"). Physical properties remain good at elevated temperatures. Excellent heavy duty bearing alloys with very good abrasion resistance and excellent resistance to repeated, severe impacts. Poor anti-seizure properties, relatively poor conformability and embeddability, hence require good reliable lubrication, hard shafts and proper shaft alignment, with both shaft and bearing machined to fine surface finishes.

Silicon Brasses Moderate-to-high strength alloys with good corrosion resistance and favorable casting properties. Used for mechanical products and pump components where combination of strength and corrosion resistance is important.

Copper Beryllium Alloys Relatively high-strength materials with good electrical and thermal conductivity. Used where bearings and bushings with a good combination of strength and conductivity are needed. Excellent heavy duty bearing alloys, but do not tolerate misalignment or dirty lubricants and generally should be used against hardened steel shafts, with both shaft and bearing machined to fine surface finish.

Leaded Copper Ultrahigh-lead alloys for special purpose bearings. Alloys have relatively low strength and poor impact properties and generally require reinforcement. Excellent for boundary-lubricated situations, or where lubrication is uncertain.

Lead-Free Bronzes Formulated to mimic the properties of the leaded alloys but without the safety concerns of lead. Friction, yield and tensile strength and corrosion resistance are all comparable to leaded bearing bronzes.

Table 2.2. Summary of Cast Bronze Characteristics.

22 Initial Selection of Candidate agricultural machinery C94100 Materials Based on Past Usage air compressor wrist pin bushings C90300 To know what has been used aircraft accessory drives C90700 for a specific application is a very aircraft control bushings C93700 powerful way to start with a given aircraft landing gear bushings C90300 material selection. Even if the aircraft-carburetor bearings C94300 application is not quite the same as what you are working on, it will automotive spindle bushings C92700 most likely form a good starting automotive transmission thrust washers C83600 point. Table 2.3 is a compilation backs for lined bearings C93800, C93700, C93600 of different bearing applications bearings for earth moving machines C92700 using cast bronze alloys. See boring bar guide bushings Al Bronzes 9 Rippel . It is suggested that you brass and copper rolling-mill neck bearings C93800 select several materials from this bridge bearings C90700, C90300 list; do not be too restrictive in the cam bearings Al bronzes type of application. Again, this cam bushings for diesel engines C93700, C93200, C93600 should not be taken as the last cam bushings for farm equipment C93700 word on materials selection, cam bushings for mechanical devices C93700 but rather the starting point. Knowledge of how well previous clutch pilot bushings C90500 designs fared can be an invaluable connecting-rod bushings for farm equipment C93700, C92700 help in such cases. cornpicker snapping roll bushings C83600 cranes and hoist bearings C94100 Another way to make an initial materials selection is by crankshaft bushings for farm equipment C93700 considering that a leaded-tin crankshaft main bearings C93600, C93700 bronze such as C93200 is the most deep-well pump-line shaft bearings C83600, C89320 commonly used material. We may deep-well pump-bowl bushings C93800, C93600 then examine this choice in light drum bushings for cranes C93800 of some other considerations drum bushings on earthmoving machinery C93800 given next. earthmoving equipment bushings Al bronzes electric-motor bushings C93700, C93200, Additional Materials Selection C94100, C93600 Considerations elevators C94100 After we have made some fire-pump bushings C94300 initial materials selection we can freight and streetcar bearings C93800 apply some additional criteria. For fuel and water-pump bushings C93700 all the different choices of bearing garden tractors C93700 materials there are two critical gas, gasoline and diesel engine bearings C93800 conditions that must be met. gasoline pump bearings C93800 These are: gear bushings for farm equipment reducers C93700, C92700 ■ the material must be strong gear bushings for motorcycles C93700 enough to withstand the peak generator and distributor bushings C93200 applied bearing stresses, and guide bushings for piston rods C93200, C89320, C93600 ■ it must have sufficient guide bushings for rams C93200, C89320, C93600 temperature capacity. guide bushings for valves C93200, C89320, C93600 guide post bushings Al bronzes heavy earthmoving equipment bushings Al bronzes

Table 2.3. Candidate Bearing Bronzes Based on Past Usage. (cont. p. 24)

23 High levels of applied bearing hydraulic glands seals C93800, C89320 stress may cause plastic flow of the hydraulic pump bushings C94300 bearing material. This should be avoided because it will otherwise king pin bushings for off-highway equipment C93800 cause severe dimensional changes. lathe bearings C93700 Similarly, high operating tempera- linkage bushings for machine tools and presses C90500, C93700 tures will cause a reduction in the locomotive bearing parts C94300, C92700, C93600 bearing strength and higher poten- low-pressure valve bearings C83600, C89320 tial for wear and seizing. machine tool bearings C94100,C90700, To obtain exact values of Al bronzes, C90300 maximum recommended stress main bearings for presses C93200 and temperatures for the different main bearings for refrigeration compressors C94300 alloys is very difficult and depends manifold bushings for earthmoving machinery C83600 a lot on the nature of the bearing marine equipment C90700 design. Typical values may be mechanical linkage bushings for farm and C93200, C93600, C89320 found in the literature of, for exam- material handling equipment ple: Rippel9, CDA7, and DIN10. mechanical linkage for farm equipment C92700 Table 2.4 lists typical maximum and packaging machinery bearing pressures and operating temperatures that can be used with motorcycle-engine bearings C93200 careful bearing design practice. outboard motor crankshaft bearings C93700 The range of temperatures and piston pin bushings C90500, C93600, pressures within each alloy class C93700 (diesels) results from the various piston pins for packaging equipment C93700 levels of soft phase present. power lawnmower bushings C93700 Materials with higher soft phase content (more lead) tend towards power shovel bushings Al bronzes the lower temperature and propeller bushings C83600 pressure levels. pump sleeves C83600, C89320 These numbers should be used rail and heavy equipment trunnion bearings C93700 as guidelines only. Also, we typically railroad car wheel bearings C93700, C93800, C94300 select a material that is just strong reduction-gear pinion bearings C94100 enough for the job. Allowance should be made for possible over- Table 2.3. (cont.) Candidate Bearing Bronzes Based on Past Usage. (cont. p. 25) load conditions. Other tribological characteristics to consider are the propensity Max. Recommended Range of Max. Recommended Operating Temp. Operating Pressure, to seizure and the resistance Bearing Alloy °C, °F MPa psi to wear. These are indicated in Leaded Red Brass 180-230 356-446 20-30 2900-4350 Table 1.1 on a relative basis. Tin Bronze and 170 338 25-35 3625-5075 This table should be consulted for Leaded Tin Bronze further refinement of the materials High-Leaded 170 338 15-25 2175-3625 selection. The relative wear Tin Bronzes resistance of a bearing material may also be judged from the Aluminum Bronze 300 572 50-70 7250-10155 identation hardness. It is well High Strength Brass 200 392 30-50 4350-7250 known that hardness plays a Copper Beryllium Alloys >200 >392 50-200 7250-29010 major role in preventing wear. It Leaded Copper 160 320 10-15 1450-2175 should be realized however that increasing hardness levels will Table 2.4. Maximum Recommended Operating Temperatures and Pressures.

24 Physical Properties of Bearing rod bushings C93800, C93600 Materials rod bushings for refrigeration compressors C94300 Other properties such as thermal roll bushings C90500 conductivity, thermal expansion roll neck bearings in rolling mills C93700, Al bronzes coefficients and density are also roller bushings for conveyors C93200 important tribological parameters, especially so when high bearing rolling mill bearings C90500, C93800, C93700 temperatures are expected. Table 2.5 seals C93800, C89320 list values for the most commonly sleeve bushings for cranes and draglines C93200, C93600 used physical properties. 10 7 spacer bushings and bearings for pumps. C93800, C89320, C93600 This data is from DIN ; CDA ; 9 speed reducers bearings C93700 and Rippel . The values listed in this table should only be used as a spindle bushings for farm equipment and trucks C93700, C93200 guideline. More specific informa- spindles and connecting rods in farm equipment C93700 tion should be obtained once a spring bushings for farm and C83600 more definitive material selection automotive equipment has been made. A range of values spring-shackle bushings C93200 suggests that slightly different results were reported by the refer- starter motor bushings C93200 ences. The range for the hardness steel mill equipment C90700, C93600 values may also reflect the differ- steering-knuckle bushings C93700 ent heat treatment conditions for supercharger bushings C94100 the alloy. textile machinery bushings C94100 thrust washers and components for chemical C93600 Compatibility with Working process equipment Fluids torque-tube bushings C93200 While for the most part we design sliding bearings using track-roller bushings for crawler tractors C93700, C93200 grease or oil as the lubricant, we trolley wheel bushings C90300 do not need to restrict ourselves to trunnion bearings C90300, C92700, C93700 these lubricants. In fact the use of turbocharger spindle bushings (floating type) C94100,C89320 sliding bearings can be turned to enormous advantage when using a turntable bushings Al bronzes given process fluid as the lubricant. valve guides C90500 The use of seals, always a weak valve rocker arm bushings C93700 link in the design, can then be avoided altogether. Typical process fluids valve stem nuts Al bronzes such as water, gasoline, kerosene, water-lubricated bushings C94300 creosote, paints, and even food wristpin bushings C93700, C93600, C89320 products such as ketchup and molasses are used as lubricants. Table 2.3. (cont.) Candidate Bearing Bronzes Based on Past Usage. Even highly corrosive fluids such as hydrochloric and muratic acids reduce the resistance to seizure, using a fairly large indenter, and can be used as liquid lubricants for and also reduce the conformability thus, averages the hardness reading sliding bearings. and embeddability. Typical hardness over a fairly large domain. The It is also important that the values are indicated in Table 2.4. indicated hardness values do not bearing and shafting materials to The values are indicated as Brinnell necessarily reflect the hardest hardness. This is a hardness test possible phase in the alloy.

25 Leaded Red Brass C83600 CuSn5Pb5Zn5 55-60 80-100 11.6-14.5 95 14 19 11 72 42 8.90 0.322 Tin Bronzes C90300 70 145 21.0 96 14 18 10 75 43 8.80 0.318 C90500 75 150 21.8 103 15 20 11 75 43 8.72 0.315 C90700 80 150 21.8 103 15 18 10 71 41 8.77 0.317 Leaded Tin Bronzes C92200 65 130 18.9 96 14 18 10 70 40 8.64 0.312 C92300 70 140 20.3 96 14 18 10 75 43 8.77 0.317 C92700 77 145 21.0 110 16 18 10 47 27 8.78 0.317 High-Leaded Tin Bronzes C93200 CuSn7Pb7Zn3 65-70 100-120 14.5-17.4 100 14.5 18 10 59 34 8.80 0.318 C93400 60 110 16.0 76 11 18 10 58 34 8.87 0.320 C93500 60 110 16.0 100 14.5 18 10 70 40 8.87 0.320 C93600 65-70 135(0.5%) 19.6 77 11 18.5 10 49 28 9.05 0.327 C93700 CuPbl0Snl0 60-70 80-110 11.6-14.5 76-90 11-13 18 10 47 27 9.00 0.325 C93800 55 110-140 11.6-20.3 72 10 18.5 10 52 30 9.25 0.334 C94300 AMS 4840 48 105 15.2 72 10 18 10 63 36 9.30 0.336 C94100 approx. CuPb20Sn5 45-50 60-80 8.7-11.6 75 10.9 19 11 59 34 9.30 0.336 High-Strength Brasses C86300 215 400-450 58.0-65.3 98 14 22 12 35 20 7.83 0.283 C86400 90 170 24.7 96 14 20 11 28 16 8.33 0.301 C86800 approx. CuZn37Mn2A12Si 150-170 280-350 40.6-50.8 100 14.5 19 11 65 38 8.10 0.293 Aluminum Bronzes C95300 140-170 — — 110 16 16 9 63 36 7.53 0.272 C95400 170-195 — — 107 16 16 9 59 34 7.45 0.269 C95500 195-230 — — 110 16 16 9 42 24 7.53 0.272 C95800 160 240(0.5%) 34.8 114 17 16 9 36 21 7.64 0.276 C95520 approx. CuA110Fe5Ni5 140-150 250-280 36.3-40.6 120 17 16 9 60 38 7.60 0.275 C95800 approx. CuA19Fe4Ni4 180-220 480-530 69.6-76.9 118 17 16 9 27 16 7.60 0.275 Silicon Brasses C87200 80 150-200 21.8-29.0 100 14.5 17 9 28 16 8.36 0.302 C87600 135 220 31.9 — — — — 28 16 8.30 0.300 Copper Beryllium Alloys C82800 180-380 500 72.5 130 18.9 22 12 123 71 8.30 0.300 Leaded Copper C98820 30-45 65 9.4 75 10.9 16 9 80 46 — — Lead-Free Bronze C89320 70 120(0.5%) 17.4 98 14.2 18 10 56 32 8.80 0.318

Table 2.5. Physical Properties of Bearing Materials. be used are compatible with the resistance ratings for various cast process media is given in CDA's process fluid. Extensive corrosion bronzes as a function of different Copper Casting Alloys8.

26 NOTES:

27 Design of Boundary Lubricated Bronze Bearings DESIGN OF BOUNDARY LUBRICATED BRONZE BEARINGS

Boundary lubrication occurs Conditions Leading to characteristics are even further with sliding interfaces where enhanced when lubricants such as the transmitted load is carried Boundary Lubrication greases are introduced into the completely by the contacting While the nature of the interac- interface. Friction and wear now asperities. While a liquid lubricant tions that take place between asper- decrease dramatically, and very may be present, an insignificant ities during sliding conditions are long lives may be obtained from portion of the load is supported impossible to predict (as of today), grease-lubricated bronze bearings. by it. Wear occurs with boundary there are several macro parameters To determine the boundary lubrication. The wear rate is depen- that force a lubricated tribological lubrication performance of several dent upon the design parameters system to operate in the asperity or commonly used cast bronze bearing (materials, load, sliding distance) and boundary mode of load bearing. alloys, the International Copper the ability of the lubricant to form These conditions are; Research Association sponsored protective, sacrificial chemical com- ■ High load extensive research that produced pounds to avoid direct metal-to- ■ Low speed invaluable data for use in the design metal contact. When significant ■ Low lubricant viscosity of such bearings. This research also chemical action takes place, the ■ Starting and stopping furthered the understanding of the role of temperature is very important. ■ Rough surfaces underlying phenomena that set the System temperatures may result 11 ■ Irregular surfaces limits of operation. Glaeser et al , from some ambient condition, or ■ Lack of suitable or adequate conducted an extensive set of lubri- they may be self-generated due to film formation mechanisms cated journal bearing wear experi- sliding under friction conditions. ments on the following ■ Inadequate clearance The level of friction will of course bearing bronzes; depend on the nature of the asperity ■ Misalignment C93200 Leaded Bronze boundary interaction and the films The above list indicates that C90500 Tin Bronze that may form here due to the almost every tribo-system will C95400 Aluminum Bronze chemical kinetics. operate at one point or another in the lubricated wear regime. There- C94500 High-Leaded Bronze. The whole aspect of boundary fore, correct design considerations It was found that the bearing lubrication is thus a complex phe- and criteria must be kept in mind in performance could be presented on nomena of chemical, thermal and the design stage of every system. a plot of sliding velocity versus the mechanical interactions. It is very mean bearing stress, or a so-called hard to predict what may happen Boundary Lubricated lubricated wear map, Figure 3.1. for a given situation, but we can This lubricated wear map for say a few things about what brings Wear Data for Cast a journal bearing shows three areas on the boundary lubrication mode. Bronze Bearings of wear identified as follows: The accompanying program, As mentioned under the mate- Region of Moderate Wear, -7 Bound, is based on the experimental rials selection of the various cast (K = 3xl0 ) data and the modeling presented in bronzes, the superior boundary Region of High Wear, -6 this chapter. lubrication aspects of copper based (K=l xl0 ) alloys are second to none. These Region of High Temperature.

29 runaway with rapidly increasing temperatures. Such behavior is indicative of at least partial hydrodynamic films separating the sliding surfaces at the lower stresses. How- ever, electrical continuity measure- ments always showed metal-to-metal contact or the absence of a complete lubricant film. At the lower speeds of 0.025 m/s (0.08 ft/s) and 0.05 m/s (0.16 ft/s), the coefficient of friction is typically 0.1 at all bearing stresses. The maximum allowable stress was determined by frictional heating, but no abrupt transition to runaway temperatures occurred. Figure 3.1. Lubricated Wear Map for a Journal Bearing. Alloy C90500 (tin bronze) shows similar behavior to Alloy The high-wear limit occurs at wear might be expected to take C95400, but has a distinctly lower a temperature of about 150 C (302 F). place is critical in the design of load capacity at 0.10 m/s (0.33 ft/s). At temperatures above this, severe bearings. This limit was determined The reduction in load capacity is 11 wear develops and failure takes by Glaeser and is summarized in directly related to the stress at place. This temperature limit will Figure 3.2 for the four alloys. which the transition from low depend on the nature of the materials Unlike the original publication by friction coefficients (0.01) to high 11 and lubricant and is thus very sys- Glaeser , which averaged data coefficients (0.1) occurs. tem specific. from four separate tests to obtain Alloy C93200 (leaded bronze) The experimental results show the wear transition limit, here the is similar to alloy C90500 at that wear and bearing distress lowest pressure limit at which a speeds of 0.10 m/s (0.33 ft/s) and becomes unacceptable at bearing transition occurred for any one 0.15 m/s (0.49 ft/s), but shows the stress levels from 24-28 MPa speed is taken as the upper stress lowest load capacity of all four (3500-4000 psi). In many grease- limit. Original comments by alloys at 0.025 m/s (0.08 ft/s) and lubricated continuous rotation Glaeser on the behavior of the 0.05 m/s (0.16 ft/s). Reduced load bearing installations, 3.5 MPa materials follow: capacity at the lower speeds results (500 psi) is accepted as a maxi- Alloy C95400 (aluminum from higher coefficients of friction mum. Therefore, boundary lubri- bronze) shows the most consistent of up to 0.18, which cause a corre- cated bearings are seldom operated and highest load-capacity results spondingly higher thermal input at bearing stress levels much above over the speed range studied. At from friction. 28 MPa (4000 psi). Some high- speeds of 0.10 m/s (0.33 ft/s) and Alloy C94500 (high-leaded strength bronzes, such as alu- 0.15 m/s (0.49 ft/s), the maximum bronze) shows the most inconsis- minum bronze, are operated at allowable operating stress is sharply tent performance of the four alloys, bearing pressures on the order of limited because of excessive tem- which results from variability in 70 MPa (10,000 psi) in oscillating peratures from frictional heating. the transition from low friction to motion applications (airframe Up to the maximum allowable high friction as load is increased. bearings, for instance). stresses at these speeds, the coeffi- This alloy also produces low coef- cient of friction is typically 0.01 or ficients of friction (0.01 to 0.02) less. Exceeding the maximum at the lower speeds as well. As a Lubricated Wear Limits results in an abrupt increase in the result, the transition from accept- for Bearing Bronzes coefficient of friction to values able to unacceptable stress levels The maximum load-speed around 0.1. This, coupled with the was fairly sharply defined at limit at which moderate lubricated higher applied load, causes a thermal 0.05 m/s (0.16 ft/s).

30 Lubricated Wear Rates for Bearing Bronzes The wear rates of four bearing alloys were measured by Glaeser at 7 MPa bearing pressure and 0.10 m/s sliding speed. The results are plotted in Figure 3.3 as the average of three bearings for each material. The four alloys are clearly divided in their extent of wear under identi- cal running conditions. Alloy C95400 was the most wear resistant, followed closely by alloy C90500. Alloy C93200 had an intermediate wear resistance, Figure 3.2. Load-Speed Limits for the Different Bronze Materials. while alloy C94500 experienced a large amount of wear. All four alloys showed the classic high initial wear rate ("break-in") followed by a leveling off to a more constant, lower value. Predictions of the amount of diametral wear encountered over a given time period for the different materials may be made by fitting the data to a wear model. For the lubricated wear data, an equation that follows the typical Archard- Holm12 works well, see Appendix A for further details on this model. The predicted amount of diametral wear is given by: Figure 3.3. Lubricated Wear Behavior for Four Bronze Materials.

Bearing Temperatures temperature, and it is in keeping where; The energy loss, W, in the with the experimental findings. d = diametral wear bearing may be calculated from The dissipated energy will give K = wear coefficient the product of frictional force and rise to a bearing temperature P = nominal bearing stress sliding velocity, viz: Fn/B•D increase. For boundary lubricated W = fFnU Fn = applied radial load bearings, the heat generated is D, B = bearing diameter and width The coefficient of friction, f, removed by air cooling of some H = hardness here should be chosen carefully. It form. This may be natural convec- depends very much on the lubrication or, in some cases, forced ts = total sliding time convection. It is a very complex U = sliding velocity tion conditions. For the grease- lubricated bearings we will assume mechanism that involves many To account for the high wear that the friction coefficient is 0.1. difficult to estimate variables. To rate in the initial break-in, two This will be an overestimate in make the model workable from an different wear coefficients are need- most cases, but it will give us a engineering point of view, we ed for this model. These are incor- conservative estimate for the bearing simplify it to the following: porated in the Bound program.

31 reflect a temperature limit, or a the various bronzes used in the where: wear limit. Also, strict reciprocity investigation. Glaeser suggests it is between pressure and speed cannot more indicative of the grease limit, heat flow to ambient be used, as is apparent from the rather than the bearing materials heat transfer coefficient bearing temperature equation. per se. Accepting the fact that the to ambient experimental limit is in fact more a outer surface area of grease limit, we may make some bearing housing Design Considerations recommendations for the maximum bearing temperature There are several critical safe temperatures in a bearing from ambient temperature considerations for the design of a the known temperature limits for successful boundary lubricated different types of greases, see Typical values for the heat bronze bearing. Adherence to these Table 3.2. transfer coefficient are calculated design considerations make it more Besides the type of filler used by a method detailed in Appendix B. likely that the bearing performance for the grease, the base lubricant By equating the heat generated to will be satisfactory. will also significantly affect perfor- that convected away to the envi- Bearing diameter is usually mance. Overall however, the tem- ronment, we get a heat balance as dictated by shaft stiffness consider- peratures shown in Table 3.2 may follows: ations and is thus not something be used as guidelines. we can truly select at random. Additional information on the Bearing width is normally chosen selection and application of greases so as to keep the width-to-diameter 13 Hence at the point of equilib- may be found in NLGI . ratio close to unity, though this will rium we obtain: be dictated somewhat by the mean Maximum Allowable Bearing projected bearing stress levels for Stress the application. Other important considerations are discussed next. The maximum allowable Note, this expression suggests bearing stress depends on the type that constant temperature lines on of material used for the bearing the lubricated wear map should be Maximum Allowable Bearing and the form of lubrication used. hyperbolas, assuming the friction Temperatures For grease-lubricated bearings and heat transfer coefficients stay Glaeser found that rapid operating in the boundary lubrica- constant. The limit of operation increases in bearing wear occur tion mode, the maximum bearing region where high wear takes place when the bearing temperatures stress is normally kept below 3.5- has the general shape of a hyperbola. exceed about 150 C (302 F). This 5 MPa (508-725 psi). From the test 11 This limit on the bearing tempera- is often accompanied by sudden data by Glaeser , it is clear that ture was found to be about 150 C increases in friction coefficients some of the materials are capable (302 F) for the four materials from low numbers like 0.01-0.03 of much higher stress levels. evaluated. to high values like 0.10-0.15. We However, it is recommended should thus calculate the theoretical to use only these higher levels under special design circumstances Bearing PV Factors temperature rise based on the higher friction coefficient. when space is critical. Examination of the wear equation and the bearing tempera- The 150 C (302 F) temperature ture equation shows that both use limit is well below the capability of the product of pressure and velocity. Grease Type Maximum Temperature, °C, °F Comments This is sometimes referred to as the Lithium soap based 140 284 General purpose "PV" factor of a bearing. Some Calcium soap 60 140 Relatively cheap design manuals use this as a general 320-392 Wide temperature range rule for boundary lubricated bearing Clay-based 160-200 design. As we see from the above Table 3.1. Recommendation for Grease/Bearing Temperature Limits. two expressions, this could either

32 Normally the projected bearing boundary lubricated bearings. The The original surface finish of the area is selected such that the shaft plays a major role in the bearing material will change during recommended stress levels are not wear-in process and subsequent the "wear-in" period early in the life exceeded. performance is directly influenced of the bearing. Some wear of the by wear-in. For the leaded bronzes, shaft will also occur owing to the Clearance cold-rolled steel, finished to 0.5-0.7 hard intermetallics in the bronze Clearance values for boundary µm (20-28 µin.) Ra (arithmetical microstructure. A shaft surface lubricated sleeve bearings should average of roughness), can be used. finish of between 0.25-0.7 µm (10- be larger than for hydrodynamic However, steel shafting, AISI 28 µin.) Ra is recommended. lubrication. This is because the 1045, hardened to HRC 30 bearing has to operate in the pres- minimum, ground and polished (or Wall Thickness and Flange ence of wear debris and any other comparable steel), is a better Dimensions adventitious debris that would choice and certainly should be used with tin bronzes containing no lead. Bronze sleeve bearings are normally be flushed out by a flow- normally contained within a housing ing lubrication system. In addition, Harder shafting is appropriate as long as it has the fracture tough- or shell. A reasonable criteria for sufficient clearance space is needed wall thickness is that the bearing, to allow injection of fresh grease ness and fatigue strength required for the application. When using when installed in its housing, into the bearing periodically. should provide adequate strength to Recommended clearance values harder shafting, it is important to keep the surface roughness levels support the imposed loads without range from 0.2% to 0.5% of the elastic or thermal distortions which journal diameter. Sufficient clear- low to prevent abrasive wear of the bearing by the shaft. would destroy the "built-in" ance should also be allowed for geometry of the bearing. possible reduction in clearance In general, increasing wall thickness is required for increasing bore diameter. Thin bearing walls and heavy housings provide more strength than the opposite arrange- ment, but heat transfer from the bearing may be impeded. If severe wear of bearing material is permissible and expected, adequate material thickness should be provided. Anticipated temperature rise is another consideration when wall thickness is specified. Large variations in temperature, and different coefficients of thermal expansion for bearing and housing materials, Figure 3.4. Recommended Clearance for Grease-Lubricated Bearings. combine to produce expansion and contraction forces that make bear- due to different thermal expansion ing dimensions and fits difficult to coefficients. Figure 3.4 shows the Surface Finish maintain. Varying clearances with- recommended diametral bearing In boundary lubricated bronze in the bearing can result. 14 clearances as suggested by Bartz . bearings, the shaft surface finish is To locate a bearing axially in a important. Since the shaft should housing bore, a shoulder flange as Shafting always be harder than the bearing shown in Figure 3.5 may be provid- The hardness and surface fin- material, steel asperities will tend ed. This can be used as a marginal ish of steel shafting is an important to plow the bronze bearing surface, thrust surface as well. consideration in the design of producing an abrasive type wear.

When a bearing is pressed or shrunk in its housing, unequal expansions can also cause stressing of both members. If the bearing material yields, cooling may change the original interference fit and result in a loose bearing. However, these difficulties may be minimized by proper design. A final requirement is that wall thickness should be at least three times the depth of any grooving. Solid bushing minimum and maximum wall thicknesses and recommended flange sizes according to ISO-4379 are shown in Figure 3.6. For flanged bushings, it is recom- Figure 3.5. Dimensions for Straight and Flanged Bushings. mended that the flange height, by housing bore. Recommended hflange, is at least equal to the bushing bearing is to press or shrink the bearing fits for general applications with wall thickness, twall. Also the in the housing with an interference fit. cast bronze bushings in cast iron or flange thickness, tflange, is made the Uniform wall thickness over the entire steel housings are H8/p7 according same as the wall thickness. bearing is easily maintained by either to ISO-1101. For aluminum hous- Maximum wall thickness and of these processes. ings, the interference may have to flange dimensions may be Standard stock bushings with be increased by as much as 0.001 D desirable for high load applications finished inside and outside diameters are to provide satisfactory tightness. where a lot of interference fit is available in sizes up to approximately For high-temperature work, the required. This will reduce the 125 mm (4.92 in.) bore size. Stock possibility of buckling. recommended interference fits bushings are commonly provided with may need to be adjusted to allow slightly over nominal OD sizes. Since for expansion differences between Methods of Retaining Bearings these tolerances are built into standard bearing and housing and to avoid Many different techniques are bushings, the amount of press fit is yielding of the bearing material. used to ensure that a bronze bear- controlled ing stays put within its housing. The method used depends upon the particular application, but it is important to ensure that the unit lends itself to convenient assembly and disassembly. One goal to keep in mind is that the bearing wall should be uniformly thick to prevent introduction of weak points in the construction which might lead to elastic or thermal distortion.

Press or Shrink Fit The most common and satisfactory technique for retaining the Figure 3.6. Recommended Wall Thickness and Flange Dimensions for Solid Cast Bronze Bushings.

Thin walled housings require into machine bores. Lubricants ■ Dowel pins somewhat lighter press fits. For should not be used when employing ■ Housing caps fabricated bearings, tolerances of shrink-fits as this may lead to These methods may be easier both bearing OD and housing bore hydraulic blister formation than a press or shrink fit, but they should be specified to produce the between the bearing and the are much less satisfactory as they recommended interference fits. housing bore. do not result in a very solid con- As a result of a press or shrink When a bearing is pressed into nection. They should only be used fit, the bore of the bearing material the housing, the driving force for light duty applications and cer- "closes in" by some amount. In should be uniformly applied to the tainly not when dynamic loads are general, this diameter decrease is end of the bearing to avoid applied to the bearing. approximately 70% to 105% of the upsetting of the bearing. A steady interference fit. Any attempt to pressing force is preferred. An accurately predict the amount of insertion arbor should always be Grease Grooving Patterns close-in, in an effort to avoid final used. Also important are the mating Because grease will only go clearance machining, should be surfaces, which must be clean, where it is pushed or dragged, it avoided. smoothly finished and free of will not distribute itself in a bearing Shrink fits may be accom- machining imperfections. like a liquid lubricant will. plished by chilling the bearing in Therefore, grooving is generally dry-ice and alcohol or in liquid Keying Methods cut into the inner surface of a bush- nitrogen. These methods are easier Many different ways are used ing. There are many grooving than heating the housing and are to fix the position of the bearing designs used. Some of the more preferred. Dry ice in alcohol has a with respect to its housing by common designs are shown in temperature of -78 C (-108 F) and "keying" the two together. Possible Figure 3.7, where the bushing has liquid nitrogen boils at -196 C (- keying methods are: been unfolded along an imaginary 321 F). The use of a hollow ■ Set screws line opposite the grease feed hole. mandrill filled with liquid nitrogen ■ Woodruff keys Most of these configurations are will greatly extend the useful ■ Bolted bearing flanges easily cut with special tools. working time to install bushings ■ Threaded bearing OD

Figure 3.7. Various Grease Redistribution and Supply Groove Patterns: (a) straight axial, (b) circular, (c) axial and circular, (d) butterfly groove, (e) single loop, (f) double loop, (g) figure eight, (h) double figure eight, (i) chain-link, (j) ladder.

35 Straight Axial Supply Groove Recirculating Configurations types of groove patterns, see Where the applied load is One difficulty with grease Figure 3.8. At a bearing stress of predominantly in one direction and lubrication is to keep the grease in 3 MPa (435 psi), the chain-link continuous rotation exists, then the the bearing without affecting the type groove consumes about 20 optimum design is to use one or load capacity greatly. Configu- times less grease than the straight more axial feed grooves, see rations such as those shown in axial groove. Figure 3.7a. The axial slot should Figures 3.7f, g, h and i, and to lesser be oriented so it is in the unloaded extent in e achieve this. These Groove Geometry area of the bearing. In addition to configurations also tend to be The grooves should not extend distributing the grease, the groove preferred for longer bearings. The out to the ends of the bearing and also provides a reservoir for the side grooves recapture the lubri- should not take up too much of grease and a trap to collect wear cant and bring it back to the grease bearing area providing load support. debris. Axial lubricant grooves may supply location for redistribution. Conventional design calls for the have an angular extent of as much Dirt and debris are however also grooves to be machined to within as 30°. If one groove is specified brought back into the load zone about 10% of the ends of the bear- this should be positioned between and this may lead to abrasive wear ing so the lubricant is kept within 90° and 120° upstream of the in the bearing. The grease applica- the bearing area. However, under direction of applied load. If two tion hole is typically placed oppo- heavy load conditions where bear- grooves are used, they should be site the point of load application ing temperature is high and debris diametrically opposite at 90° to the with these configurations. accumulation becomes a problem, applied load. Axial groove length Configurations such as shown it is better to have the grooves exit should be approximately 0.8 times in Figures 3.7f and h reduce the at the ends of the bearing so peri- the bearing width. load capacity by virtue of interrupt- odic flushing of the bearing with ing the load bearing area. This fresh grease will remove the debris Circumferential Grooves reduction is less for g and i. Both g and lubricant decomposition and i are also very efficient in products collected in the grooves. If the applied load varies recapturing the grease. Tests con- Experiments with "non-recirculat- considerably in direction, then the ducted by Bethmann and reported ing" grooves indicate lower wear choice of a central circumferential 15 by Spiegel show the benefit in than with recirculating grooves. groove is preferred, see Figures 3.7b recirculation of grease for these and c. Such a design has, however, a lower load capacity than an axial groove bearing of equivalent size. Typically the width of the circumferential groove is approximately 20% of the total bearing width. When using this groove pattern for oscillating bearings, it is preferred that the axial groove spacing is about the same as the sweep during a given oscillatory move. This is to ensure that the entire shaft is relu- bricated for each cycle. If rotational motion is very limited, then the ladder-groove as shown in Figure 3.7j should be used.

Figure 3.8. Experimental Grease Consumption for Different Groove Patterns (Sliding speed is 1 m/s, bearing diameter is 80mm).

36 There are many different cross-sectional groove geometries. The most commonly used are shown in Figure 3.9. It is very important that the groove edges are all properly rounded. Sharp edges tend to act as a grease scraper and thus remove grease from the shaft rather than apply it. Recommendations for groove depth and width for grease lubrication are given by Rippel9, and are shown in Figure 3.10.

Grease Consumption Figure 3.9. Common Forms of Grease Grooving. From Rippel9. Low wear rates can be maintained only over long periods of Besides bearing size, we must and where; time when grease is present in the also expect that it is a function of B = bearing width, bearing and can perform its intended bearing speed, bearing loading, U = sliding velocity, function. Because grease is lost operating temperature and groove Λ = loading factor, and from the bearing due to side leak- patterns. These effects were incor- G= environmental factor. age, and also somewhat consumed porated in an empirical expression This expression does not take due to chemical action, we have to by Henningsmeyer as reported the groove pattern into account, make provisions for its replace- 14 by Bartz that may be used for and we will thus need experimental ment. This may be done on a con- a more refined estimate. refinement to get better estimates. tinual basis or as is more common, Henningsmeyer gives the following This can easily be done through on a periodic basis. Certainly a expression for the grease consump- very careful monitoring of the major advantage of grease-lubricat- tion rate: bearing temperature. Rapidly ed cast bronze bearings is that they increasing temperatures suggest are not overly critical for a continual grease supply, and can thus be used more grease is needed. for periodic relubrication. Grease consumption for grease- lubricated bearings is a very complex phenomenon, and thus far has been treated on an empirical basis only. It clearly must be a function of the bearing size, and practical experience indicates that it lies somewhere between:

Vf = (40-1000) Aj, where Aj is the inside bearing area (pBD). This expression shows a very wide range of possible grease consumption, and can only be used as a starting point.

Figure 3.10. Recommended Groove Dimensions for Grease Lubrication.

37 Regreasinq Intervals Thus, the regreasing interval The bearing loading factor, L, Assuming that grease leaks may be estimated from; is a function of the bearing load from a bearing at a constant rate and the bearing operating tempera- during operation, we may estimate ture. The following expression is a regreasing interval. This interval suggested: is based on the assumption that the maximum allowable grease deple- When using this estimate, it tion is 25% of the total grease should be realized that the estimate volume in the bearing (excluding is for the actual operating time and the supply and redistribution not clock time. Also grease quanti- P = bearing stress, 3 grooves). The total grease volume ties smaller than about 0.04 cm 2B = bearing temperature. in the bearing is given by; (.002 in.3) are difficult to deliver.

When using this expression care should be taken that L>1. The environmental factor may where: be estimated from the conditions B = bearing width, surrounding the bearing. Values are D = journal diameter, suggested in Table 3.2. CD = diametral clearance.

Likely ingression of dirt or dust G = 0.01 Likely ingression of water G = 0.02 Likely ingression of dirt and water G = 0.03

Table 3.2. Values for Estimating the Environmental Factor of Conditions Surrounding the Bearing.

38 NOTES:

39 Hydrodynamic Journal Bearing Design HYDRODYNAMIC JOURNAL BEARING DESIGN

When the load on a bearing is fully supported by hydraulic pressure in a lubricant film, there will be a finite separation between the bearing and shaft. The two components do not contact each other and wear is nonexistent or very low. If this separating lubricant film is solely caused by the motion of the journal surface relative to the bearing, we speak of hydrodynamic lubrication. In a hydrodynamic journal bearing, the part that moves is generally the journal (as with the shaft of a motor), sometimes the bearing (for example, the wheel of a Figure 4.1. Basic Journal Bearing Component Designation. Conestoga wagon), and sometimes both parts (such as a connecting monly used. The lubricant to a applied to the journal. Because this rod that joins a piston and the journal bearing may be supplied by film consists of lubricant, it may crankshaft in an automobile engine). a non-pressurized lubricant feed be sheared as much as necessary Since many motors, engines such as an oil ring, a wick or other without any permanent damage and other machines incorporate means. It may also be supplied by to the bearing. hydrodynamic journal bearings, the pressure. Pressure-fed systems may Pressure in the lubricant film annual production of hydrodynamic include supply and return lines for develops when we drag lubricant journal bearings is in the billions. the oil, a storage tank, filter, and into a converging space, A, shown Not only is the hydrodynamic temperature and pressure regula- in Figure 4.2a. This drag is impart- journal bearing very common, it is tors. Similar to grease-lubricated ed on the fluid because fluids like superior to rolling-element bearings bearings, grooves may be used to to adhere to surfaces. As the lubri- for many purposes because it can spread the oil along the bearing. cant is drawn around into the con- carry heavier loads, operate at verging space, a certain amount of higher speeds, is less expensive, the fluid is forced out against the is more reliable, and can last Load Capacity of viscous action, B. This results in a indefinitely when designed well. Hydrodynamic Journal local pressure. As the gap con- A typical hydrodynamic jour- Bearings verges, the sideways squeezing nal bearing is shown in Figure 4.1. becomes more and more difficult Because the hydrodynamic The key to hydrodynamic and higher pressures develop. In operation is dependent on the lubrication is the formation of a the end, only a small but finite presence of lubricant, we typically lubricant film that separates the amount goes through the smallest require a constant supply. Several shaft from the bearing. Pressure in gap in the bearing. After the small- ways of achieving this are com- this film, when integrated over the est gap is passed, C, the space bearing area, will support the load

41 From the foregoing discussion of action in a hydrodynamic journal we might expect that the load supported by the film would be a function of the viscosity, speed, journal area, and radial clearance. These variables are often combined into a convenient dimensionless variable called the Sommerfeld number, in honor of the work done by Sommerfeld in finding a closed form solution to load capacity. Hence the Sommerfeld number is given as:

Figure 4.2a. Lubricant Flow in a Figure 4.2b. Pressure Distribution Journal Bearing. Resulting from the This number contains all the Lubricant Flow. essentials required to find the load a given bearing can support. It is a opens up again, and the pressure small, then a fairly large gap at C is function of the bearing eccentricity rapidly drops to ambient. possible because not much pressure and the bearing width-to-diameter The resulting pressure distribu- is needed to support the load. For ratio, as shown in Figure 4.3. In tion around the bearing looks thus high loads, this gap will be small Figure 4.3 the independent vari- something like that shown in because higher pressures are needed. able is given as the eccentricity Figure 4.2b. Note, this pressure is Full-film formation occurs ratio. This ratio is defined as: generated by the action of the as long as the smallest gap in the lubricant film in the bearing, bearing is greater than the com- and not by the supply pressure. bined roughness on the journal and Integrating this generated pressure bearing. The smallest gap in the over the bearing area results in the bearing occurs along the line of and forms yet another convenient applied load. centers in the vicinity of the load. dimensionless group to use. We The beauty of a journal bear- The line of centers makes an angle, may use the eccentricity ratio to ing is that we can get this hydrody- f, with this load line. The actual calculate the minimum film namic action almost for free. The value of the minimum film thick- thickness in the bearing by; only special provisions we need to ness is given as hmin. From the make are to ensure a suitable clear- displacement of the journal center ance exists between the shaft and relative to the bearing center, see the bearing surface and sufficient Figure 4.1, and the radial clear- The calculation of Sommerfeld lubricant is supplied at all times. ance between the journal and the numbers has been studied and The clearance is easy to achieve, bearing, we calculate this gap as; researched at great length ever simply by making the shaft some- since Osborne Reynolds formulated what smaller than the bearing bore. the differential equations to solve The rest of the action takes place for this problem. For in depth by itself. The thickness of the gap, where C = radial clearance details of the mathematics for example, adjusts itself to the (Db-Dj)/2, and involved, the reader is referred to load that is applied. If this load is e = bearing eccentricity.

42 Cameron16, Constantinescu17, Szeri18, Hamrock19, and others. Tables prepared by CD A20 contain some of the earliest calculated data by Raimondi21. Besides the load capacity we also need data on the required oil flow and the friction generated by the bearing. These data may be found in CDA20. The data from these tables were incorporated into the Hydro program for bearing calculation purposes.

Thermal Effects Heat generated by lubricant shear in a bearing is of major importance in practical applications, mostly because it causes an increase in lubricant temperature. For most lubricants, the viscosity Figure 4.3. Sommerfeld Number as a Function of the Eccentricity Ratio for decreases with increasing tempera- Bearings of Different Width. ture. This directly affects the load capacity of a bearing as may be Cast bronzes are excellent Transition to Hydrodynamic choices as bearing materials seen from the Sommerfeld number. Lubrication Lower viscosities mean lower load because of their very high heat transfer coefficients. Hence struc- For a given journal bearing capacity for a fixed Sommerfeld with a fixed R/C ratio, the coeffi- number. tural cooling of the bearing is quite cient of friction may be plotted as a A second effect of increasing feasible for a moderately loaded function of the hydrodynamic bearing temperatures is the hydrodynamic bearing. parameter, ηN/P. This friction is due increase or decrease in bearing Heat removal by a continuous to the viscous shear of the lubricant clearance. If materials with different supply of cooler lubricant is also a only. In real bearings however the thermal expansion coefficients are very effective way to cool the bear- surfaces have a definite roughness, used, the resulting clearance ing. Both structural and lubricant and it may be expected that at high increase can result in a significant cooling of the bearing are incorpo- Hydro values of the eccentricity, i.e., loss of load capacity. Both the rated in the program. viscosity-temperature and thermal Additional cooling of the bear- when the film thickness is small, expansion effects are incorporated ing may be obtained by supplying some interaction of asperities on into Hydro. lubricant at higher than required opposing surfaces occurs. When Frictional heat can be removed flow rates through the bearing. In the eccentricity goes to unity, the by conduction through the bearing fact automotive engine bearings friction coefficient plot needs to be components or by oil flow. In some are cooled this way. The current modified to include the boundary Hydro cases, sources external to the bear- version of does not have layer friction, see Figure 4.4. ing bring additional heat into the this feature. It is expected that a Minimum friction occurs at an system so that the lubricant and future release will have this option, eccentricity of ε ≈ 1- hc/C. The bearing then act as a cooling together with some additional transition takes place where the film

system as well. lubricant supply methods. thickness is equal to the convection heat transfer coefficient, hc.

43 For any practical application, we always operate to the right of this point. This region is known as the hydrodynamic friction region, and the slope reflects the operating viscosity. The point at which the action is mostly hydrodynamic is called the transition point. Notice that viscosity has an impact on this transition. During start-up from cold conditions, the viscosity is typically high. Hence a lower transition speed is required to get hydrody- Figure 4.4. Modified Bearing Friction due to Boundary Lubrication. namic action. When the bearing comes to a halt after extensive to make sure the bearing will not components are under load. Based operation at higher temperatures, suffer a scuffing failure. It may be on this aim, it would seem that we viscosity is much lower, and the required to adjust the surface should try to make the separation danger of scuffing is higher. If the roughness, viscosity or bearing size between the two bearing members transition occurs at a higher speed, to make possible a reliable transi- as large as possible. This however more heat is produced at this point. tion-to-boundary lubrication. would take a lot of power, since a Critical in the prevention of scuff- hydrodynamic bearing system is ing is the choice of bearing materi- Design Criteria for Journal nothing more than a viscous pump, als. Here the cast bearing bronzes Bearings and a very leaky one at that. excel because of their very good To arrive at some reasonable boundary friction characteristics. Having considered some of the choices about the separation, we (See Chapter III.) theoretical aspects of hydrodynamic tend to make the bearing size as The transition journal speed, lubrication, we now have come to small as necessary, and at the same N may be calculated from the the hardest part of all. By what time have a reasonably safe separa- T criteria do we judge a bearing minimum recommended safe film tion. Because bearings of large design to be satisfactory? thickness and the Sommerfeld length tend to suffer more from number. Consider the critical We need to define several edge contact due to shaft deflec- eccentricity ratio as follows; design criteria. For example: what tions and misalignment, we sel- is the minimum safe film thickness domly use B/D ratios larger than 1. for operation? What are the factors Also we will try to maximize the for safety? What are the allowable safe temperature increase in the temperatures? In some cases, you bearing, within the limits as speci- may have internal guidelines to fol- fied. A design that follows these low based on previous experience. guidelines will be safe and, at the It may also be very worthwhile to same time, not waste unnecessary use Hydro and reverse engineer a energy. We extract the critical speed couple of designs that you know to from this as: give satisfactory performance. Minimum Recommended Film Thickness Bearing Size From the above discussion, it The design of a hydrodynamic is quite clear we need to establish This speed can now be used in the bearing is aimed at avoiding exten- some criteria for the separation of boundary lubrication calculations sive metal-to-metal contact when

44 the two surfaces. This suggested separation will depend on a number of judgmental variables that should be carefully considered before making a final choice. Consideration should be given to the following: ■ size of the bearing (larger bearings require larger separation just due to manufacturing tolerances), ■ applied loads during running and during starting periods, ■ bearing deflections and frame distortions, ■ consequences of a bearing failure, loss of life, etc. With these considerations in Figure 4.5. Minimum Recommended Film Thickness for Journal Bearings. mind, we now must choose a minimum acceptable film thickness for the bearing. Below are some guidelines based on empirical information and with a solid experimental basis. They should help to minimize the probability of any trouble. ■ Minimum recommended film

thickness h2min. In most cases, this is a function of machine size and surface roughness. In the absence of any internal company guidelines, Figure 4.5 may be used. These data are based on marine bearing Table 4.1. DINx Guidelines for Minimum Allowable Film Thickness for design practice, and usually Journal Bearings, μm. have some form of load sharing, see Hill22. Design Clearances bearings, use the values as 10 given in Figure 4.6. The DIN guidelines for the Equally important in journal ■ Seizure concerns: When the minimum allowable film thickness bearings is the correct choice of bearing and journal materials take the operating speed into bearing clearance. We have seen have different thermal expan- account. This is a more refined that small bearing clearances result sion coefficients we must method. in high operating temperatures. make sure this effect is Comparison between the values On the other hand, large clearances considered. Also we need to in Figure 4.5, and those in Table 4.1 result in loss of performance. check and see if the chosen indicates they are roughly in line The right clearance is thus clearances can result in bearing with each other. very important. seizure at temperature ■ Eccentricity ratios of less than ■ Clearance ratios: for small extremes. This may be the 0.7 should be avoided because bearings use a clearance ratio case at either end of the of possible dynamic instabilities. 500 ≤ R/C ≤1000, with the larger clearances for high temperature spectrum. If at speed application. For larger all possible, the thermal

45 ture rise, Δθ (from lubricant inlet to exit), is from 30-50 C (86-122 F), depending on the materials used. It is restricted to keep thermal gradients and subsequent distortions to reasonable levels. Because of the high thermal conductivity of cast bronze materials we can use the upper limit.

Bearing Stresses and Materials ■ Satisfactory boundary lubrication must be provided for start-up purposes. This translates into a maximum bearing pressure of about Figure 4.6. Minimum Recommended Diametral Clearances for Journal Bearings. F/LB < 2.5-3 MPa (362-435 psi) at start-up. When running, this expansion coefficients of the criteria is mostly for oil oxida- may be considerably higher. If bearing should always be tion life. This gets rapidly the start-up load results in equal to or greater than the shorter as the operating pressures higher than those journal material. This temperature exceeds 80 C recommended we should requirement is automatically (176 F). There are two differ- provide a hydrostatic assist. satisfied for most common ent cases that are considered; ■ Maximum recommended configurations of steel jour- • Bearings with oil circulation: bearing pressure calculated

nals and cast bronze bearings. θmax is 100-125 C (212-257 F). over the projected area It is also possible to get start- • Self-contained bearings (air depends on the bearing up seizures when very quick cooling): θmax is 90 C (194 F). materials chosen. For cast heating of the shaft allows it When synthetic lubricants are bronze bearings, we will to grow more rapidly than used we add 30-40 C (86-104 F). follow the guidelines as the bearing. Seizures of this indicated in Table 2.3, page 24. type typically occur very fast ■ Permissible lubricant tempera- (within 10 seconds of start-up). To check for possible seizure conditions, perform a layout calculation as shown in Figure 4.7. The tolerance range on the nominal bearing and shaft size is taken into consideration.

Maximum Temperatures and Temperature Rises ■ Maximum permissible

lubricant temperature θmax This is calculated at the cavita- tion point in the bearing. This

Figure 4.7. Clearance Variation with Temperature for a Bearing with a Higher Thermal Expansion Coefficient than the Journal 46 Safety Factors some cases it may be necessary journal is fixed? What if the load One has to be very careful with to provide special construc- rotates relative to the bearing? We the use of safety factors in hydro- tional modifications to reduce will provide some simple rules for dynamic bearing design. Liberal effects of distortions. this in the next section. We will use of large safety factors will lead also provide some simple guide- to very inefficient bearing design, Lubricant Grooving lines on easy ways to avoid severe and in some cases, may even lead Grooves may be machined into edge-loading problems. to design failures. the bearing surfaces to promote the ■ Safety factor on load. In distribution of lubricant. The appli- Bearing and Load Rotation practice, a safety factor of 1.2 cation and precautions are similar So far the load capacity and to 2.0 on load is used depending to grease grooving, see page 35. friction analysis for journal bear- on how well the load is known. For oil-lubricated bearing grooves, ings has dealt with the simple case Under no circumstances should use only about 60% of the suggested of journal rotation. In many practical a factor greater than 2.0 be used. grease groove width. applications we also have bearing ■ Safety factors on oil supply rotation, load rotation, or combina- should be about 1.5 to 2.0. Machining Considerations tions of all three. It is quite simple to extend the foregoing analysis to ■ Cylindricality and waviness include these additional possible Lubricants and Filtration errors should be less than 50% rotations by using an equivalent ■ In modern designs the viscosity of the smallest expected film journal speed. Note that all rota- grade does not exceed ISO thickness for the operation of tions are referenced to the load VG100 (lighter oils are the bearing, thus: vector. The general case of journal, often used). waviness error < 0.5 x h 2min bearing and load rotations may be ■ Adequate filtration must be ■ The maximum surface rough- analyzed by considering journal provided. Typical filtration ness should be less that 25 and bearing rotations independent levels should be for particles to 40 times the minimum from load rotations, see Figure 4.8. greater than 50% of the mini- separation, or: mum operating film thickness. rms surface roughness In the first case, the equivalent speed results from the addition of Filtration below 10 μm may

be about 30-40% of the largest Ne = Nj + Nb - 2 NF surface dimension to avoid Other Journal Bearing where: load based deflections that Details N = equivalent journal rotation might affect performance. e The analysis in Hydro is for Nj = original journal rotation ■ Shaft and housing deflections the journal bearing with the journal N = bearing rotation should be kept to levels less b rotating and the bearing stationary. NF = load vector rotation. that half of the minimum film What if the bearing rotates and the thickness in the bearing. In

47 extremely important. In the following discussion of each of these topics we will refer to the mating surface as the shaft. This is only out of convenience, and the details of the discussion apply equally well to the mating surface of a thrust bearing, a spherical bearing or any other form of bearing surface.

Shafting Materials The predominant number of bearing applications involve steel as one of the members. The predominant shafting material is steel, either a carbon steel, tool steel, gray or nodular cast iron. Stainless steel shafting may be used, but care must be taken in selecting grades that have high seizure resistance when used with certain bearing alloys. Fatigue strength of the shaft is important if cyclic loading is anticipated on the bearing. The use of the case-hardened shafts in many machine elements is very common. Hardened steel is very resistant to both adhesive and abrasive wear. This wear resistance Figure 4.8. Equivalent Journal Rotations for Different Journal, Load and increases with hardness. From a Bearing Rotations. practical standpoint, the maximum obtainable hardness on carbon steel As might be apparent by now, The last column in Figure 4.9 components is about 65 HRC. At the sense of the rotations should be shows some very simple construc- this level of hardness, the ductility kept in mind. We have used clock- tions that often can easily be incor- of the steel is very low, and failure wise here as positive. (Note: Some porated in a new design. Often by brittle fracture can occur very of these bearings will require a ring these 'silent' features introduce rapidly if the shafts are overloaded. groove supply in order to function a significant robustness into To avoid this problem we use case- correctly.) the design. hardening techniques. Geometries to Mitigate Edge In case-hardening, the surface Mating Surfaces of the material is hardened to maxi- Loading due to Shaft Bending The choice and properties of mize wear resistance, while the core Edge loading due to shaft the mating surface in a bearing is kept soft to avoid brittle fracture bending is a serious problem, and system are just as important as the of the part. This hardened case may can significantly reduce load selection of the bearing material. be achieved on shafts by suitable capacity. In Figure 4.9 are several Specifically, the material type, the induction heat treating of moderate simple ideas that can reduce the surface roughness and preparation to high carbon steels or by carbur- severity of the problem (from methods, and the hardness of the izing of a low-carbon based metal. Bartz14). mating bearing surface are

48 softer bearing material by the shaft, its roughness should be low. Typical nominal Ra values are around 0.2- 0.3 μm (8-12μin.) for the shaft and double this value for bearings. If the differential hardness between bearing and shafting materials exceeds 150-200 HV, the shaft roughness should be made correspondingly lower. Also, as the hardness of the bearing material increases, its roughness and the shaft roughness should be made correspondingly lower to avoid two-body abrasion by the bearing material on the shaft and by the shaft on the bearing.

Hardness of Shafts The surface hardness of the shaft should be high to resist wear by the bearing and to resist wear by the ever-present abrasives such as silica. In most cases, the shaft is Figure 4.9. Methods to Avoid Edge-Loading Problems. the expensive, difficult-to-change The latter method is used most patterns on the shaft is highly member which increases the commonly for mass production undesirable. importance of minimizing wear techniques and is less subject on the shaft. to error. Tin bronzes require harder Surface Roughness Levels steel shafting and a good finish. The best levels of roughness Surface Preparation The hard constituents in tin on the bearings and shaft are often bronzes will tend to wear in the The most common method of a compromise between the cost to steel surface and improve its finish. surface preparation is grinding. produce these levels of roughness, Aluminum bronzes require a hard- This may be done by centerless and the amount of run-in wear that ened steel shaft or chromium grinding or by conventional methods. can be tolerated. Wear of the bearing plating on the shafting. Copper The control of surface roughness and shaft will lead to larger clear- beryllium alloys also require a and waviness levels, surface lay ances, and thus part of the wear life hardened shaft material—prefer- and fuzz generation are very of a bearing system may be con- ably a tool steel. important in the satisfactory sumed simply by not producing the As a good starting point, we operation of a bearing. optimum roughness on the new should aim to make the shaft at To keep circumferential wavi- components. least 150-200 HV points harder ness, out-of-round and grinding Wear in a bearing system can than the hardest constituent in a chatter under control, it is suggested take the form of adhesive wear, bearing material. This will ensure that roughness and waviness traces abrasive wear or a combination of that almost all the wear in a bearing be made periodically. Chatter should both. We control adhesive wear system will take place on the softer be kept to a minimum, and lobing through a proper materials selec- component. If the hardness differ- effects of grinding must be within tion, while abrasive wear can be ential is greater than this, care the circumferential waviness limits. controlled through roughness. To should be taken to lower the levels The presence of 3 to 7 definite lobe avoid two-body abrasion of the of roughness for the shaft. We tend

49 high temperatures, excess wear debris in the bearing and eventual seizure. 2. Grooved shaft produces same conditions as 1. 3. Bearing thrust face larger than mating surface produces same conditions as 1. 4. Opposing oil inlet holes within one bearing bushing prevent proper oil flow across bearing surface. The pressure gradient produced results in high temperature, excessive wear and eventual seizure. 5. Dead end acts as a reservoir until filled and then prevents proper oil flow across bearing surface, resulting in high temperatures and excessive wear. (A point to remember is that it is just as critical in a bushing application to get Figure 4.10. Common Design Problems with Bearings. the oil out as it is to get the oil in.) to rely on a certain amount of avoided in order that satisfactory 6. Fillet ride also prevents oil polishing of the shaft by the bear- bearing life may be realized. flow. ing material. This becomes more The index numbers shown on 7. Two precision bushings in and more difficult as the shaft the figure are explained as follows: one short hole can result in gets harder. 1. Shaft ending within the misalignment due to accu- bushing allows journal to mulated manufacturing wear a matching ridge in the Other Issues tolerances. This forces the relatively soft bearing mater- Designing a good bearing can shaft out of line and results ial. This restricts free flow of be helped immeasurably by knowing in edge loading. Do not use the lubricant across the the pitfalls of bad design. Figure 4.10 coaxial flats on shaft if they heavily loaded area, causes shows some of the common design can be avoided. edge loading and results in errors presented by Pesek23 to be

50 NOTES:

51 Computer Programs COMPUTER PROGRAMS

The CDA computer programs In Figure 5.1 maximum loaded at, say, 10 MPa, 10 cm/s Bound and Hydro provide a means allowable bearing pressure (load (1500 psi, 20 fpm) or to investigate for the design of either boundary divided by the projected bearing low friction hydrodynamic design or hydrodynamic lubricated jour- area) is plotted versus journal for a relatively slow speed, lightly nal bearings. These programs can speed in rpm, for a 25 mm (1 in.) loaded bearing, say, 3.5 MPa, be used on either IBM PC compat- diameter bearing. In the hydrody- 15 cm/s (500 psi, 30 fpm) that ible computers or Macintosh. They namic portion, the maximum oper- had been originally considered a will accept units in both English ating bearing pressures are shown boundary lubricated bearing. (IPS) or Metric (SI). for two different grades of oil. In The hydrodynamic program the boundary lubrication zone, Background to the Programs Hydro develops a table of film maximum bearing pressures are The hydrodynamic bearing thickness, eccentricity ratio, power shown for two different friction program is based on the material absorption, bearing oil temperature coefficients presented in the previous chapter. and oil flow rate for a given range The boundary lubrication pro- The bearing performance charts of of radial clearances and at a fixed gram is based on grease lubrication Raimondi and Boyd21, developed load. The designer inputs the load, only. The criterion for maximum from solutions to Reynold's equa- speed (rpm), bearing size, inlet bearing pressure is grease thermal tions have been computerized. The temperature and selects a lubricant. instability. Boundary lubricated program automatically iterates the The program stores a database of bearings wear continuously during converging estimates for oil tem- properties for various types of oils service, and bronze bearing alloys perature. This is essentially a great- and inserts these data as needed. have maximum allowable com- ly expanded computerized version The boundary lubricated bear- pressive loads. Thus it is possible 9 of the Rippel manual. ing design program Bound is for to consider grease-lubricated oper- The boundary lubrication heavily loaded, slow speed grease- ation for a hydrodynamic designed program is based on empirical data lubricated bearings. It develops a journal bearing that is clearly over- table that includes estimated wear, maximum operating bearing temperature (with or without air cooling) and adsorbed horsepower for four classes of bronze bearing alloys. The program also sets bounds for acceptable operating temperatures. Both programs cover a wide range of bearing operating conditions, from high bearing pressures and low speed (25 MPa, 5 cm/s (3500 psi, 10 fpm)) to fully hydrodynamic conditions (14 MPa, 5 m/s (2000 psi, 1000 fpm)). Both programs overlap in the range of operating conditions as shown in Figure 5.1. Figure 5.1. Maximum Bearing Stress Operating Range for a 25 mm (1 in.) Journal Bearing for Both Boundary and Hydrodynamic Lubrication. 53 for grease-lubricated bearings and to provide quick estimates for heat-transfer equations with the bearing performance for the design Macintosh On the desktop: maximum operating temperature engineer. Special effects such as Double click on floppy disk icon as a criterion for design, see the influence of grooving, edge Double click on file "HYDRO" Chapter III. loading or non-circularity of the The user should be aware that components is not taken into account IBM PC DOS both programs have certain limita- by these programs. Nevertheless, At the DOS prompt: tions. Specific assumptions and they do provide a good first cut Type "HYDRO" limitations are: estimate of bearing performance. Press RETURN/ENTER ■ Programs are for steady state For such exceptional running loading only, conditions as turbulent flow or Figure 5.2. Opening the Program. ■ Program are for full 360° bearing instability (whirl), other moved to a directory on your hard- bearings only, analytical approaches should be drive. Make sure you also copy the LUBE.TRU file to the same directory. ■ Oil is supplied at a single oil used. In any instance, it is best to All the input and output files will supply hole located opposite run bearing tests to verify the then show up on this directory. the applied load, design before freezing the design. Something like the message ■ The effect of oil oversupply is Graphical presentation of the in Figure 5.3 should appear. not considered in the cooling hydrodynamic program output is Answer the query. If you know or performance of the bearing, an option. For IBM DOS compatible of a file on the current disk that ■ The designer should ensure systems this requires a VGA or relates to this program enter it. If suitable alignment to avoid higher level monitor. Hard-copies you want to start afresh, answer N, edge loading of the bearing, of the graphs can be obtained on dot matrix printers that are Epson or simply press RETURN because ■ The heat transfer model for the default value is (N)o. both Hydro and Bound compatible, or on laser printers that are HP Laserjet compatible. The screen display shown in assume certain boundary Figure 5.4 will appear. Graphics can be viewed, saved, conditions (See Appendix B This table allows you to input for more details.), and printed on all Macintosh systems. bearing load, speed, size, ambient ■ Installed clearances are at temperature, bearing thermal prop- room temperature, 20 °C (68 F), erties, lubricant and lubricant inlet ■ Clearances indicated are Using the Hydrodynamic temperature. radial clearances, Program The table appears with the val- ■ The included viscosity data on Insert the floppy disk and, ues inserted from a default set or SAE oils is for high viscosity depending on the system you are the last run. Some of the values are index oils.(See Appendix C using, follow the instructions in default values such as thermal con- for details.) Figure 5.2. ductivity and expansion (values The two programs are intended NOTE: The programs may also be assumed for steel and bronze).

CDA HYDRODYNAMIC LUBRICATED BEARING CALCULATION

Rev= 16.2 Date 930722 Time 12:13:43 Project: Journall Units: IPS

Do you want to get an input file from disc? Y/N (N)

Figure 5.3. Opening Screen.

54 *********************************************** ***********************

CDA HYDRODYNAMIC LUBRICATED BEARING CALCULATION

Rev= 16.2 Date 930722 Time 12:13:43 Project: Journal 1 Units: IPS

*********************************************** ***********************

ITEM # NAME UNITS VALUE

A Bearing Radial Load (lbf) 500 B Shaft Speed (rpm) 3500 C Bearing Inside Diameter (inch) 1.0000 D Effective Bearing Width (inch) 1.0000 L U Total Shaft Length (inch) 4. 00 F Housing Outside Diameter (inch) 3.00 G Housing Width (inch) 1.00 H Shaft Thermal Conductivity (Btu/h ft F) 29. 00 I Housing Thermal Conductivity (Btu/h ft F) 28. 00 J Ambient Air Temperature (F) 83 K Air Velocity (fpm) .0 L Shaft Therm. Exp. Coeff. (mu-in/in F) 6.30 M Bearing Therm Exp. Coeff. (mu-in/in F) 10. 20 N Lubricant Name SAE 30 0 Lubricant Supply Temperature (F) 83 P Heat Remvl by lube (On/Off) On Q Heat Remvl by amb.air (On/Off) Off

U Toggle the Unit system

Enter Letter for Item # or Y(es) to accept ? ()

Figure 5.4. Input Table. Initially, the program assumes your (heat removal by air) on. In addi- significant process in bearing system uses an oil supply that the tion the air velocity may be supplied performance, it is essential that the bearing pumps through its clear- (line K) if forced air cooling is correct conductivity and expansion ance, removing the bearing heat. anticipated. Appendix A has more values are entered in the program NOTE: Clearance values in this details on the assumptions made in for materials selected for a given program are radial. the thermal modeling and some of design. Therefore items H,I, L and the different ways that control can M should be adjusted for materials. Entering Bearing Data be exercised over the heat transfer The input values shown above mechanisms. Lubricant Selection can be changed by selecting their The bearing and shaft material To select a different lubricant, item number (such as B or b for thermal properties (thermal expan- press N and a table of lubricant shaft speed) and then entering the sion and thermal conductivity) are names will appear. When you appropriate values. If the system parameters used in the program. select a lubricant, its temperature- uses a drip feed or oil vapor sys- These values in the default table viscosity properties are entered tem, air cooling can be assumed by (File Journal 1) are for C93200 into the program. Note that when selecting line Q and toggling line P bronze and AISI 1040 carbon steel. you select the lubricant, the Since thermal expansion is a 55 program displays the lubricant name However, if you do not save as matrix or HP Laserjet compatibility and asks if it is correct. If you answer a file, the same input values will of your hardware.) Y(es), you are returned to the input come up again when you activate Two plots are available: mini- table. the input table. While the program mum film thickness versus radial The default data in the data file is computing, a series of messages load for all the clearance values is based upon high viscosity index show the percentage of completion displayed in the output table and (HVI) lubricants. It should be real- and information related to the bear- minimum film thickness versus ized that there may be a significant ing design. The results then appear mean film temperature for all the variation in viscosity-temperature as a table. See Figure 5.5. clearances. Both plots show the behavior from one nominally same NOTE: On some systems you may selected operating load, recom- lubricant to another. Hence it is be asked to press Return half way mended minimum film thickness always best to enter your own mea- through the output. and whirl instability threshold. sured viscosity data for the specific The output reiterates the input Examples of the graphics are lubricant you intend to use. values, including the lubricant shown in the Figures 5.6 and 5.7. Additional lubricant data may selected and its properties. The table Figure 5.6 shows minimum be added to the data file as shown of calculated data below the input film thickness versus radial load in Appendix C. table shows a number of perfor- for 10 different clearances. The mance values for 10 different radi- design load is shown as the applied Entering Data in Alternate Unit al clearances. load line. A vertical dot-dash line Systems (NOTE: If you are using diame- shows recommended minimum You may use IPS (inch, lb- tral clearances in your design, rather film thickness limit. A dashed line force, seconds) units or SI (meters, than radial clearance, multiply the shows the limit for reduced load or Newtons, seconds) units in your clearance values in the table by 2.) "whirl limit." This limit represents design. The unit system can be The output table provides a condition where the journal posi- changed at any time during data power loss in the bearing, minimum tion has come close to the bearing input by toggling between the two oil flow rate, mean oil film temper- axis. This can lead to instability when pressing U. You can mix ature, minimum oil film thickness, (vibration) or shaft whirl. your units as they are entered. For eccentricity ratio and a rating for Figure 5.6 shows another oper- instance, if you have entered load, the reliability of the calculations ating characteristic for a hydrody- shaft speed and bearing dimensions (confidence). It also prints out the namic bearing: shearing of the in IPS units (lb , rpm, inches) and minimum recommended film lubricant film in the bearing gener- f thickness and clearance in two you have the shaft and bearing ates heat. The program estimates statements below the table. thermal expansion in SI units the bearing temperature expected After the bearing calculation (µm/m °C), you can toggle to SI for the input parameters. It is wise results are displayed, the program units by pressing U and Return. to operate at as low a bearing gives choices for G(raphic) display The table converts to all SI units. temperature as possible to minimize of performance characteristics or You can then enter your thermal thermal breakdown of the lubricant. sending the output to a F(ile). expansion values. If you wish to Note that temperatures decrease Graphics provides a valuable chart- have results in IPS, press U and with increasing clearance and ing of bearing behavior over a Return. The table is then in IPS units. increase with increasing film thick- range of clearances and show ness. Thus with a given clearance, Calculating Bearing limitations such as recommended a fluctuating load will result in Performance minimum film thickness and bear- variable bearing temperature as the ing instability threshold. If a D(ot) film thickness varies with load. When the correct values are matrix or a L(aser) printer is on Instinctively, one would like to entered into the table, the program line, you may choose to print out choose design parameters that can be asked to compute the bearing the graphics. Otherwise it may be result in the largest value of mini- performance values by answering displayed on the screen. (NOTE: mum film thickness. However, Y(es) to the inquiry. The input can For IBM PC systems this requires other factors must be considered. be saved in a file as instructed by VGA graphics and EPSON dot the program.

56 CDA HYDRODYNAMIC LUBRICATED BEARING CALCULATION Rev= 16.2 Date 930722 Time 12:13:43 Project: Journall Units: IPS

INPUT DATA: Bearing Radial Load (lbf) 500 Shaft Speed (Rpm) 3500 Bearing Inside Diameter (inch) 1.0000 Effective Bearing Width (inch) 1.0000 Total Shaft Length (inch) 4.00 Housing Outside Diameter (inch) 3.00 Housing Width (inch) 1.00 Shaft Thermal Conductivity (Btu/h ft F) 29.00 Housing Thermal Conductivity (Btu/h ft F) 28.00 Ambient Air Temperature (F) 83 Air Velocity (fpm) .0 Shaft Therm. Exp. Coeff. (mu -in/in F) 6.30 Bearing Therm Exp. Coeff. (mu-in/in F) 10.20 Lubricant Name SAE 30 Lubricant Supply Temperature (F) 83 Heat Remvl by lube (On/Off) On Heat Remvl by amb.air (On/Off) Off

LUBRICANT DATA Name = SAE 30 Dynamic Viscosity 1.200e+l Pa.s at - 18 C Dynamic Viscosity 9.000e-2 Pa.s at 40 C Dynamic Viscosity 9.500e-3 Pa.s at 99 C Specific Heat 1.850e+3 Joules/kg C at 20 C Lubricant Density 8.850e+2 kg/mA3 at 20 C

CALCULATED DATA: PREDICTED HYDRODYNAMIC PERFORMANCE FOR DIFFERENT CLEARANCES:

Rad Clear. Load Pwr Loss Oil Flow Oil Temp hmin Ecc. Confdnc (mu-in) (Ibf) (hp) (Gpm) (F) (mu-in)

2666.67 500.0 .181 .077 98 687.67 .748 Good 2133.33 500.0 .181 .058 103 671.69 .695 Good 1706.67 500.0 .178 .044 109 635.74 .644 Good 1365.33 500.0 .172 .033 117 586.85 .598 Good 1092.27 500.0 .165 .026 125 532.61 .557 Good 873.81 500.0 .156 .020 134 478.89 .522 Good 699.05 500.0 .147 .016 143 429.45 .492 Good 559.24 500.0 .138 .013 153 385.95 .467 Good 447.39 500.0 .130 .011 162 349.27 .447 Good 357.91 500.0 .123 .009 171 318.88 .429 Good

Minimum recommended Film thickness 210.0 (mu-in) Minimum recommended Radial Clearance 1051.5 (mu-in)

Figure 5.5. Calculation Screen. 57 on bearing performance by determining minimum film thicknesses for the maximum and minimum clearance resulting from dimensional tolerances. A good approach to clearance selection involves picking the smallest clearance on the basis of the above considerations so that as the bearing wears during its life (during start-up and shutdown), the increase in clearance will stay at the lower end of the recommended values. So one might select a radial clearance of 1092 (µinch (28 µm) from Figures 5.6 and 5.7. If one should choose to use the lower clearance values included as dotted lines in the figure, one must make sure to prescribe the needed precision in manufacture and the fine surface finish values for reliable operation. The recommended clearance range in the program output is based on several investi- gators findings, from journal bearing tests. See Fuller24. Figure 5.6. Diagram of Bearing Operating Characteristics Showing Minimum Film Thickness Limit. (Data points are clearances, µinches) Air-Cooled Design Clearance during start-up. If the bearing under considera- The dotted lines in Figure 5.7 The latter condition, thermal tion is expected to operate under are radial clearances that may be gradient, becomes quite significant limited lubrication flow (such as too small for safe operation. as bearing sizes and speeds drip feed, vapor lubrication, or Selecting the optimum clearance increase. During start-up, the hydrodynamic grease lubrication), for a journal bearing requires frictional heating causes the journal the design process should be modi- consideration of several options. to expand while the bearing may fied for air-cooling only. To change Referring to Figure 5.7, it can be expand inward until the bearing the program to air-cooling mode, seen that bearing stiffness (spring housing equilibrates in temperature. press P and answer the instruction rate) increases with decreasing Thus the clearance closes down by typing "off, then press Q and clearance. during this initial phase of opera- enter "on". When the input table tion. If the clearance is too small, shows "Heat removal by lube... off" Keep in mind, manufacturing the bearing clamps down on the and "Heat removal by amb. air...on", practices limit the precision possi- journal and seizure occurs. press K to input the cooling air ble in producing a bearing. This velocity selected for your design. includes bearing I.D. and journal Therefore clearance must be O.D. tolerances, geometry (out of large enough to get through this The output shown below will round condition), misalignment, initial critical start-up period. In have different values from the and thermal gradients existing addition, one can determine the output for oil-cooled design. effect of manufacturing tolerances Generally, when air-cooling is

58 used, bearing load capacity will be reduced as compared with oil- cooling mode. To illustrate the difference in bearing operating characteristics for air-cooled and oil-cooled operation, two bearing designs will be examined. The graphic results for the two conditions are shown in Figures 5.4 and.7.

Applying the Hydrodynamic Program Results The bearing design and lubricant selection is aimed at long life and reliable operation. These concerns are of great importance when reviewing the program output. Let us refer to the graphics, Figures 5.6 and 5.7. Figure 5.2, showing minimum film thickness curves for ten clearances and a large load range, is probably the most useful chart. First of all, where the load line crosses the recommended minimum film thickness line, it can be seen that the design is within safe Figure 5.7. Mean Film Temperature versus Minimum Film Thickness. operating limits for all clearances (Data points are clearances, µinches) shown. Referring to the output table, it is Figure 5.6. An operating bearing Looking at the family of also evident that the larger clear- may be subjected to periods of load curves, it can be seen that as clearances lead to higher oil flow and changes, increased temperatures ance increases, the slope of the higher power losses in the bearing. and speed changes. The design load-film thickness curve decreases. As bearings wear in service, must be able to operate within that This means that the smaller the the original clearance will open up. set of operating conditions and still clearance, the stiffer the bearing, Therefore it is best to base the be within the safe operating limits. i.e., if the load increases, the jour- design on the minimum clearance Referring to Figure 5.7, note nal center will move a smaller or the left side of the "recommend- that with increasing clearance, the amount away from the bearing ed design envelope" shown in mean film temperature of the bear- center. In addition, the smaller the Figure 5.6. ing decreases. Heat is detrimental clearance, the closer to the recom- For the usual surface finish to oil since extended exposure to mended minimum film thickness achieved in manufacturing, the elevated temperatures causes dete- the bearing will operate. Owing to clearance should be set at between rioration. Therefore, it is wise to precision limits in manufacturing, 0.10% and 0.15% of the journal try to set the design for minimum a minimum clearance should be diameter. This limit is shown in operating temperature. Higher specified. Also, a maximum clear- Figure 5.6. Thus the "safe" operat- clearance limits might thus be ance should also be specified to ing limits of the bearing are con- advisable in a design that appears insure a reasonably stiff bearing. tained within the box outlined in to involve elevated temperatures.

59 of a file on the current disk that relates to this program enter the file name. If you wish to start afresh, answer N(o) and press return. The input table in Figure 5.12 of the appears on the opposite page.

Entering Bearing Data The above table shown in Figure 5.12 allows you to input bearing load, speed, size, ambient temperature, bearing thermal properties and bearing friction coefficient. The table appears with the values inserted from a default set of the last run. The program assumes that frictional heat is dissipated by convection, conduction and radiation. Other cooling accepted is forced air moved over the bearing housing. The program allows one to input cooling air velocity. Friction The default friction coefficient is 0.10. This is a maximum level determined in the bearing perfor- Figure 5.8. Operating Characteristics for Air-Cooled Conditions. mance experiments. It is also con- (Data points are clearances, µinches) sidered a likely starting value for Using the Boundary Macintosh newly installed bearings. As the Lubrication Program On the desktop: bearings wear in, the friction level Double click on floppy disk icon can decrease. The lower end of the NOTE: This program is Double click on file "BOUND" friction range is assumed to be based on the use of grease as 0.05. Friction coefficient levels the lubricant. IBM PC DOS tend toward the maximum at high- The Bound program is At the DOS prompt: loads, low-velocity conditions. basically for slow moving, heavily Type "BOUND" loaded grease-lubricated bronze Press RETURN/ENTER Units bearings. To initiate the program, Figure 5.10. Opening the Program. The input can be in IPS or SI insert the CDA floppy disk and, units. Inputting U will toggle from depending on the computer system The message shown in Figure one units system to the other. you are using, follow the instruc- 5.11 should appear: tions in Figure 5.10. Answer the query. If you know

CDA BOUNDARY LUBRICATED BEARING PERFORMANCE PROGRAM Rev= 6.2 Date 930725 Time 16:06:22 Project: None Units: IPS ****************************************************************************** Do you want to get an input file from disc? Y/N (N) Figure 5.11. Opening Screen. 60 The unit system can be changed during data input. For instance if you have input load in lbf, speed in rpm and bearing size in inches and you have thermal conductivity data in W/m °C, press U and RETURN and all entries in the input table will be changed to SI units. You may then enter your conductivity value. If you want your results in IPS, toggle your table back to IPS by pressing U and Return.

Calculating Bearing Performance When the correct values have been entered into the table, the program can be asked to compute the estimated operating temperature and wear by answering Y(es) to the inquiry. In boundary lubricated bearings, some metal-to-metal contact occurs, and the bearing will wear continuously during its operating life. Heat developed from friction in the bearing is not dissipated by lubricant flow. Therefore, Figure 5.9. Operating Characteristics for Oil-Cooled Operation. bearing temperature is the critical (Data points are clearances, µinches)

CDA BOUNDARY LUBRICATED BEARING PERFORMANCE PROGRAM Rev = 6.2 Date 930725 Time 16:06:22 Project: None Units: IPS

ITEM # NAME UNITS VALUE

A Bearing Radial Load (lbf) 225 B Shaft Speed (rpm) 300 C Required Operational Life (hr) 1000 D Bearing Inside Diameter (inch) 1.000 E Shaft Length (inch) 4.00 F Effective Bearing Width (inch) 1.000 G Shaft Thermal Conductivity (Btu/h ft F) 29.00 H Housing Outside Diameter (inch) 3 .00 I Housing Width (inch) 1.00 J Housing Conductivity (Btu/h ft F) 28.00 K Ambient Air Temperature (F) 75 L Air Velocity (fpm) 300.0 M Bearing Friction Coefficient (-) . 10 U Change the Unit system Enter Letter for Item # or Y(es) to accept ? () Figure 5.12. Input Table. 61 parameter for judging performance. diametral clearance caused by wear the leaded bronzes, it would be The program assumes that a bearing for four different grades of bronze appropriate to select a lower coeffi- temperature above 300 F (150 C) bearing materials. The wear is relat- cient, say, 0.08. With that change in will cause unacceptable lubricant ed to the total number of hours of input and the program recalculated, deterioration for conventional operational life. Note that the wear the results shown in Figure 5.13 mineral oil-based greases. values differ among the four bronzes. are obtained. The results shown in Figure 5.13 This may influence your selection Note that the estimated bearing will appear when the query is of material. For instance, the leaded temperature now is 43 F lower answered and Return is pressed. bronze materials (C93200 and (289 F-246 F). This would insure The output provides values for C94500) tend to operate at a lower longer lubricant life. The wear of sliding velocity, bearing stress (load/ friction coefficient than the non- leaded bronze remains the same. If projected area) and horsepower leaded bronzes. Thus one might the bearing temperature for natural absorbed by the bearing. It also want to compromise on wear life convection exceeds 150 C (300 F) shows estimated bearing tempera- for a lower operating temperature. and the value for forced convection ture with or without forced-air con- The default coefficient of fric- stays below that temperature, the vection cooling. Finally it shows tion (0.10) is conservative for leaded following message will be added a table of estimated increase in bronzes. Thus, if one selects one of to the output:

CDA BOUNDARY LUBRICATED BEARING PERFORMANCE PROGRAM Rev= 6.2 Date 930725 Time 16:06:22 Project: None Units: IPS

INPUT DATA: Bearing Radial Load (Ibf) 225 Shaft Speed (rpm) 300 Required Operational Life (hr) 1000 Bearing Inside Diameter (inch) 1.000 Shaft Length (inch) 4.00 Effective Bearing Width (inch) 1.000 Shaft Thermal Conductivity (Btu/h ft F) 29.00 Housing Outside Diameter (inch) 3.00 Housing Width (inch) 1.00 Housing Conductivity (Btu/h ft F) 28.00 Ambient Air Temperature (F) 75 Air Velocity (fpm) 300.0 Bearing Friction Coefficient (-) .10 CALCULATED DATA: Sliding Speed = 78.54 (fpm) Bearing Stress = 225 (psi) Power Loss = .054 (hp) Bearing Temperature due to Natural Convection 289 (F) Bearing Temperature due to Forced Convection 199 (F) PREDICTED DIAMETRAL WEAR FOR DIFFERENT MATERIALS: Diametral MATERIAL Wear (inch) C93200 (Bronze 660) .0458 C90500 (Bronze 62) .0114 C95400 (AL. Bronze) .0034 C94500 (Bronze 520) .0567 Figure 5.13. Results Screen. 62 "The bearing temperature exceeds results in temperatures above 150 C Of course, other means can the recommended value for use (300 F), the following message be taken to reduce the operating with simple natural convection will appear: temperature. One can increase the cooling. You must use the forced "The bearing temperature exceeds housing surface area (increase the convection cooling." the recommended value. It can be housing diameter) and/or the bear- If both the natural convection reduced by increasing the forced ing length. Bearing load or bearing and forced convection cooling convection cooling." rpm can be reduced also, if the design permits.

CDA BOUNDARY LUBRICATED BEARING PERFORMANCE PROGRAM

Rev= 6.2 Date 930725 Time 17:30:41 Project: None Units: IPS

INPUT DATA:

Bearing Radial Load (Ibf) 225 Shaft Speed (rpm) 300 Required Operational Life (hr) 1000 Bearing Inside Diameter (inch) 1.000 Shaft Length (inch) 4.00 Effective Bearing Width (inch) 1.000 Shaft Thermal Conductivity (Btu/h ft F) 29.00 Housing Outside Diameter (inch) 3.00 Housing Width (inch) 1.00 Housing Conductivity (Btu/h ft F) 28.00 Ambient Air Temperature (F) 75 Air Velocity (fpm) 300.0 Bearing Friction Coefficient (-) .08

CALCULATED DATA:

Sliding Speed = 78.54 (fpm) Bearing Stress = 225 (psi) Power Loss = .043 (hp)

Bearing Temperature due to Natural Convection 246 (F) Bearing Temperature due to Forced Convection 174 (F)

PREDICTED DIAMETRAL WEAR FOR DIFFERENT MATERIALS:

Diametral MATERIAL Wear (inch)

C93200 (Bronze 660) .0458 C90500 (Bronze 62) .0114 C95400 (AL. Bronze) .0034 C94500 (Bronze 520) .0567 Figure 5.14. Results Screen. 63 References and Appendices REFERENCES AND APPENDICES

References

1. Snow 8. CDA 15. Spiegel Snow, Richard F., "Bearing up Copper Casting Alloys, Copper Spiegel, K., Fricke, J., Meis, K.R., Nobly," American Heritage of Development Association, 1994. and Sonntag, F., Zur Verbesserung Invention and Technology, Vol 4, der Berechnungsmoeglichkeiten Number 1, 1988. 9. Rippel fuer fettgeschmierte Radialgleitlager., Rippel, H. C, "Cast Bronze Tribology und Schmierungstechnik, 2. Cameron Bearing Design Manual," Cast Vol 6, 1995. Beauchamp Tower Centenary Bronze Bearing Institute, Inc, Lecture, Proceedings of the Chicago, 1959. 16. Cameron Institute of Mechanical Engineers, Cameron, A. Basic Lubrication 1979, Vol 193, No 25. 10. DIN Theory, Longman, 1971. DIN-Taschenbuch 198, "Gleitlager 3. Rabinowicz 2,Werkstoffe, Prüfung, 17. Constantinescu Rabinowicz, Ernest., Friction and Berechnung, Begriffe," Constantinescu, Virgiliu., et al, Wear of Materials., John Wiley & BeuthVerlag, 1991. Sliding Bearings, Allerton Press, Sons, Second edition, 1995. New York, 1985. 11. Glaeser 4. Lansdown Glaeser, W.A., Dufrane, K.A. & 18. Szeri Lansdown, A.R, and Price, A.L., Buchnor, W.R., Final Report on Szeri, A. Z., Tribology. Friction Materials to resist Wear., Evaluation of Load Capacity of Lubrication, and Wear, McGraw Pergamon Press, 1986. Cast Bronze Bearings, Hill., 1980. International Copper Research 5. Pratt Association, Inc., 1976. 19. Hamrock Pratt, G.C., "Materials for Plain Hamrock, B.J., Fundamentals of Bearings," International 12. Archard Fluid Film Lubrication, NASA Metallurgical Reviews, 1973, Vol Archard, J.F., Contact and Rubbing Reference Publication 1255, 1991. 18, pp 62-88. of Flat Surfaces, J. Appl. Phys., Vol 24, 1953, p 981. 20. CDA 6. Glaeser Cast Bronze Hydrodynamic Sleeve Glaeser, William A., Materials for 13. NLGI Bearing Performance Tables, Tribology, Tribology Series 20, NLGI Lubricating Grease Guide. Copper Development Association, Elsevier, 1992. Second edition. National Grease 1969. Lubricating Institute, Kansas City, 7. CDA Mo., 1989. 21. Raimondi Copper Alloy Bearing Materials, Raimondi, A.A., & Boyd, J., "A Design and Application Guide, 14. Bartz Solution for the Finite Journal CDA TN-45, 1991. Bartz, W J., Gleitlagertechnik Teil 2. Bearing and its Application to Band 163, Expert Verlag, 1986. Analysis and Design III," ASLE Trans. Vol l, No l, 1958.

65 22. Hill For Further Reading Hill, A., "Modern Bearing Design and Practice," Transactions from Gunther the Institute of Marine Engineers, Gunter, Edgar J., Dynamic Vol 88, 1976. Stability of Rotor-bearing Systems, NASA SP-113, 1966, U.S. 23. Pesek Government Printing Office, Pesek, Laddie, and Smith, Walter, Washington D.C. 20402. "Sleeve Bearing and Thrust Washer Design Trends," SAE Transactions Lang 680426, 1968. Lang, O.R., and Steinhilper, W, Gleitlager, Springer Verlag, 1978. 24. Fuller Fuller, Dudley D., Theory and Moes Practice of lubrication for engi- Moes & Bosma, "Design Charts neers, Second Edition, John Wiley for Optimum Bearing & Sons, 1984, pp 285-287. Configurations", Trans. ASME April 1971, JOLT, pp. 302-305. 25. Holm Holm, R., Electric Contacts Pinkus Handbook, Springer-Verlag, Berlin, Pinkus and Sternlicht, Theory of 1946. Hydrodynamic Lubrication, McGraw-Hill, 1961. 26. Herschel (1919) Herschel, W. H.: Standardization of Welsh the Saybolt Universal Viscometer. Welsh, R.J., Plain Bearing Design Bur. Standards, Technical Paper 112, Handbook, Butterworths, London, (1919). 1983.

66 Appendix A. Quantifying where, where, Lubricated Wear V = volume of material worn K = wear coefficient Ar = real area of contact The above wear equation was Wear in the presence of a lubri- L = total sliding distance developed by Archard12, drawing cant occurs commonly in mecha- α = thermal expansion coefficient on Holm's25 theory of wear. The nisms and is a primary cause of constant, K, is often called the deteriorated performance and failure. The extent of wear occurring wear coefficient. It is extremely depends upon the lubrication important to realize the nature of regime encountered. In some the assumptions that have gone applications that are specifically into the derivation of this equation. designed to develop a hydrodynamic If we in anyway alter the contact film, the lubricated wear phase is conditions during asperity interac- short and transitional during run- tions, we generally expect that to It is also well known that the ning-in and during start-stop cycles alter the removal rate of material. only. When conditions of full real area of contact between two Hence the wear coefficient is very hydrodynamic film generation can- surfaces is directly proportional to dependent on the particular condi- not be established, it is the main the applied load. This is regardless tions that prevail at the asperities. mode of lubrication. It should be of whether the asperity contact was Of particular importance are the pointed out here that even subtle elastic or plastic. If the contacts are following: surface features such as machining predominantly plastic then we find • Adhesion conditions of the waviness can contribute significant that the real area of contact is materials. This in turn depends on: hydrodynamic action in what inversely proportional to the hard- • Local temperatures (depen- appears to be an otherwise non- ness of the softer of the two bodies dent upon ambient tempera- hydrodynamic application. in contact. Material removal from ture, friction and sliding the surface is thus expected to be velocity) The extremely complicated proportional to the applied load, nature of wear virtually precludes • Local cleanliness conditions and inversely proportional to hard- accurate quantitative relationships • The solubilities of the two ness. Hence the relationship to predict wear. However, for engi- materials between wear volume V and neering purposes, relationships • Asperity contact conditions: distance L, load F , and hardness exist for assessing the general wear n • Load and geometry result in H is expected to be: regime and the likely effect of elastic or plastic contact stress changing one of the variables on • What portion of the applied the resulting wear rate. load is carried by the asperities. Hence the above wear equation with a constant, K, is only valid Archard-Holm Equation where, over a range of loads and distance From very simple basic argu- V = wear volume when asperity conditions stay the ments, it may be expected that the Fn = normal load same. Experience verifies this. removal of material between two L = sliding distance The lubricated wear data surfaces in contact is proportional H = hardness obtained by Glaeser" was observed to their real area of contact. Also If the contact conditions at the to behave according to the Archard material removal may be expected asperities stays the same while model. The experimental wear to be proportional to the total dis- material is removed, then we may coefficients are as indicated in tance through which the surfaces introduce a constant of proportion- Figure 3.1. Because the bearings move. Thus it is perfectly reason- ality K, and the wear equation now were grease-lubricated, the wear able to expect that wear is: reads: mechanism remained the same during the tests, with the exception of the initial wear due to running-in.

67 The experimental data fitted to We may also calculate the The cumulative damage con- this model is used in the Bound wear depth when the bearing cept may also be used to calculate program. operates at constant speed for a an equivalent wear period for an given time t as: application with variable loading Other Forms of the Archard and speed. The time to reach the Equation same wear depth may be found from manipulation of the above It is often convenient to write equation as: the wear equation in a somewhat different form. Sometimes we Lubricated Wear under Variable prefer to express wear as a linear Loading Conditions. dimension change. For example, In many applications, the load- we may be more interested in a ing on a bearing may be variable, change of bearing clearance over and the speed may change, or the time. Such an expression may be direction of rotation may even derived from the above, if we con- reverse. To predict wear in such sider the wear volume V to be the instances, we need to apply cumu- product of the apparent area of lative damage models. Fortunately contact A and the depth of wear. this is fairly simple for bearing This allows us to derive an expres- wear calculations, provided that the sion for the depth of wear simply wear mode stays the same over the by dividing by the apparent area of life prediction period. With grease- contact, or: and oil-lubricated cast bronze bearings, this is expected to be the case. is average pressure, and To apply cumulative damage methods for lubricated wear calculations, we simply make a summa- where, tion of the product of load and A = apparent area of contact sliding distance over the prediction d = depth of wear lifetime. Hence the wear depth is the average speed. P = nominal contact pressure equation becomes: This simple model will be bearing stress very useful when using the Bound program for applications with variable loading.

We have here used the absolute The nominal contact pressure value of the product of pressure in a bearing is thus simply the radial and velocity because wear is essen- load divided by the projected bear- tially path dependent. For situa- ing area. This is taken to be D*B. tions where the load and sliding To calculate the time t to reach velocities are essentially constant a critical wear depth d*, when slid- over long periods of time, and ing takes place at constant velocity where a distinct finite number of U, can be done by letting L = U t. these constant usage periods exists, Rearranging gives: we may write:

KPU where, U = constant sliding velocity, d* = critical depth of wear 68 Appendix B. Thermal General Heat Transfer Model 5. The radiative emmissivity Modeling of the Bearing A general heat transfer model coefficients for the shaft and of the bearing is shown in Figure bearing housing have been In the performance analysis of B.1. This model allows for the gen- fixed at 0.8 and 0.7, journal bearings, the operating eration of heat within the bearing, respectively. temperature plays an important and for dissipation of this heat by 6. The model assumes that the role. For the boundary lubrication convective, conductive and radia- shaft is solid. mechanism, the peak temperature tive heat transfer. This model is With these assumptions and at the interface is the key to deter- essentially implemented in both restrictions in mind, the program mine the life and suitability of the programs with the following calculates the heat transfer from construction. For hydrodynamically restrictions: the bearing by conductive, convec- lubricated bearings, it is the strong 1. Shaft diameters are the same tive and radiative mechanisms. All temperature dependence of viscosity of this heat has to be absorbed by for lubricating fluids that interacts on both sides and are the same as the bearing inside the ambient air. Additional heat with the load capacity. Hence it is transfer from the bearing is essential that reasonably accurate diameter. 2. Both sides of the shaft have allowed for by the convective prediction of the bearing tempera- mechanism in the lubricant flow. ture be performed. This prediction, an equal length of exposure. The important dimensions however, has to be in keeping with 3. The ambient air temperature required by the program for the the requirements of the solution. is the same for the shaft and the bearing. heat transfer calculations are Computer programs have been 4. The effective length of the shown in Figure B.2. developed which are capable of shaft is at such a location obtaining a thermo-hydrodynamic where the shaft temperature solution to the energy and Reynold's Control of Heat Transfer is ambient. This means equation for a given bearing con- Mechanisms there is no heat flux in the figuration. The difficulty which In designing a bearing applica- shaft at that location. often remains here is what bound- tion it is often required that the ary values to use for the various bearing performance be checked component temperatures. Thermal analysis of the entire bearing construction for different boundary conditions is then required. An alternate solution, and the one used here, is one whereby a reasonably simple empirical model is used for the heat loss from a bearing configuration, and then to use the resulting temperatures as a way to calculate an effective viscosity in the bearing. This method essentially is a film energy balance, whereby the frictional heat generated is carried away by conductive and convective mechanisms. The simplicity of the model, however, necessitates some restrictions in the applicability of the model. These restrictions are explained in Figure B.1. General Model of the Various Heat Transfer Modes from a the next section. Journal Bearing.

69 and forced-air convection. By changing the "air velocity," we can alter the degree of forced convection. When the "air velocity" is set to 0, only natural convection and radiative heat transfer from the housing take place. The heat transfer from the shaft is, however, based on the forced convection mode due to shaft rotation. Because bearings must have a mechanism whereby viscous/frictional heat is removed, a number of safeguards have been implemented. Thus it is not possible to set Figure B.2. Definition of Bearing Dimensions for the Heat Transfer Model. both "heat removal by lubricant" and ambient air to "Off." The pro- under a number of different condi- turned "Off," only convection by grams will reset the value for tions. This is especially important the lubricant is permitted. For "heat removal by lubricant" to for bearings where the heat transfer Bound, it cannot be changed. "On." For the Bound program, the conditions are not well known. The ■ Heat removed by the lubri- heat transfer mechanism by the bearing design programs Hydro cant. Can be switched "On" or lubricant does not exist and is thus and Bound have a number of con- "Off " in Hydro. When it is not provided as an option. trols that can be exercised switched "Off," it simulates the It is recommended that the by the user to determine their heat transfer from a grease-lubri- user implement a cautious approach influence. cated bearing operating in the to the heat management of bear- These are: hydrodynamic mode. ings, i.e., always estimate on the ■ Heat removed by the ■ Control of the air velocity. conservative side. ambient air. For Hydro, this may Heat transfer from the bearing and be turned "On" or "Off." When shafts is based upon both natural

70 Appendix C. Fluid Viscosity own measured viscosity data temperature order. The for the specific lubricant you temperatures need to be Database intend to use; expressed in °C, and the The Hydro program as sup- ■ It is imperative that the user viscosity needs to be expressed plied comes with a limited fluids verifies the assumed tempera- in the dynamic units of Pa·s database. The current fluids in the ture-viscosity and the actual (a section dealing with viscosity database are: temperature viscosity behavior conversion factors is given ■ ISO VG32 of the fluid he selects. later); ■ SAE 10W ■ ISO VG46 Adding Additional Fluids to the Dynamic ■ SAE 10 Database Temperature, °C Viscosity, Pa·s ■ SAE 20 0 12.1 Additional fluids may be 100 0.013 ■ ISO VG68 added to the database by using any 200 0.00022 ■ SAE 30 word processing program. The data ■ ISO VG100 are contained in an ASCII text file ■ SAE 40 called LUBE.TRU. This file must ■ Fluid density data point. One data point on the fluid density ■ ISO VG150 be called up as an ASCII text file in kg/ms at 20°C is required. ■ SAE 50 and saved as an ASCII text file. Be (1260kg/m3); ■ ISO VG220 aware that some word processing ■ Specific heat capacity data ■ SAE 60 programs add special characters to point. One data point for the ■ Di-2 Ethylhexyl Sebacate a file if you do not save it as ASCII. specific heat capacity at constant ■ TetraChloroDiphenyl The data essential to the hydrody- volume (Cv) in Joule/kg•°K at ■ Polydimethyl Siloxane namic program are the following: ■ 293 K is required. (2350 ■ Glycerol Fluid name, ■ Joule/kg•°K). ■ Tolulene 3 temperature-viscosity points, ■ ■ Castor Oil Fluid density data point, and ■ Viscosity Conversions ■ Methanol Fluid specific heat data point. Traditionally viscosity was ■ Ethanol This information needs to be entered on a single line in comma measured by noting the time ■ Water separated value (CSV) format. A required for a given amount of The database contains the fluid to be drained by gravity temperature-related properties of typical entry, for example, for Glycerol is given below. Details of through a small orifice. The Saybolt these fluids. Because the standards Universal Seconds (SUS), the for both the ISO VG and the SAE each entry are given next. Glycerol, 0,12.1, 100,0.013, Redwood Seconds (RS), and the grades DO NOT specify the tem- 200, 0.00022, 1260, 2350 Degree Engler (E) are all based on perature viscosity behavior of these this method. The information thus ■ Fluid name. Any name up to specific grades, certain assumptions obtained is a measure of the kine- were made in selecting data for 15 characters long that describes the fluid. (Glycerol); matic viscosity. This measured these lubricants. The user should be time in seconds may be converted ■ Temperature-viscosity points: aware of and consider the following: to proper kinematic viscosity These points are used to inter- ■ The default data for mineral by the following formula due to polate the fluid viscosity at oil-based lubricant is based Herschel26: upon high viscosity index different temperature forms. (HVI) lubricants; The values must be given in the Temp 1, Vis 1, Temp 2, ■ There may be a significant Vis 2, Temp 3, Vis 3 order. Try variation in viscosity-tempera- where, ture behavior from one nomi- to use the maximum spread in temperatures for best results. It n = kinematic viscosity in nally-the-same lubricant to centistokes (cSt) another; is also recommended that you enter the data in increasing t = measured runout time. ■ It is always best to enter your

71 The constants A and B depend on the type of instrument used and are NOTES: given below.

Type of Viscometer A B Saybolt (SUS) 0.22 180 Redwood (RS) 0.26 171 Engler (E) 0.147 374

The units of kinematic viscosity in SI units are m2/s. Formerly the kinematic viscosity was quoted in Stokes or Centistokes. Conversion factors from other units of kinematic viscosity are:

Kinematic Viscosity Unit Conversion Factor 1 Stoke (St) 10-4 m2/s 1 Centistoke (cSt) 10-6 1 square foot per hour 0.258 x l0-4 1 square foot per second 929.03 x l0-4 m2/s 1 square inch per hour 0.179 x l0-6 m2/s 1 square inch per second 6.452 x 10-4 m2/s

Temperature changes affect the oil kinematic viscosity in two ways: by changing the flow resistance or dynamic viscosity and by changing the density. For hydrodynamic lubrication, we only require the change in flow resistance and therefore the dynamic viscosity. In the SI system of units, the dynamic viscosity has units of:

The dynamic viscosity of a lubricant is obtained by multiplying the kinematic viscosity by the density of that fluid. Note that the density needs to be expressed in kg/m3 and the kinematic viscosity in m2/s for this calculation. (Units are: m2 s-1 x kg m-3 = kg m-1 s-1 = N s m-2 = Pa·s) Conversion factors from other dynamic viscosity units are:

Viscosity Unit Conversion Factor 1 poise (P) 0.1 Pa·s 1 centipoise (cP) 0.001 Pa·s

1 poundforce second per square inch (Reyn) 6896 Pa·s

1 poundforce second per square foot 47.88 Pa·s 1 poundal second per square foot 1.488 Pa·s 1 pound per foot-second 1.488 Pa·s 1 slug per foot-second 47.88 Pa·s